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CQEffilGHT OEPOSm 



Compressed Air Plant 



THE PRODUCTION, TRANSMISSION AND 

USE OF COMPRESSED AIR, WITH 

SPECIAL REFERENCE TO 

MINE SERVICE 



BY 

ROBERT PEELE 

M ... 

Mining Engineer and Professor of Mining in the School of Mines, Columbia University 



FOURTH EDITION, REVISED AND ENLARGED 
TOTAL ISSUE SIX THOUSAND 



NEW YORK 

JOHN WILEY & SONS, Inc. 

London: CHAPMAN & HALL, Limited 
1920 



7/20 






^ca'^ 



Copyright, 190S, 1910, 1918, 1920 

BY 

ROBERT PEELE 



.^ -O-Zo/^ 




PRESS or 

BRAUNWORTH & CO. 

BOOK MANU^ACTURERB 

BROOKLYN. N. V. 



§)CI,A605026 

DEC 17 iS20 



PREFACE TO THE FOURTH EDITION 



In this edition a new chapter has been added: Chapter 
XXVII, on " Measurement of Air Consumption." It contains 
data on the flow of air through orifices and short tubes, and a 
discussion of different appKances for measuring air at both low 
and high pressure, including simple tank apparatus that can 
be readily improvised. 

A note has been appended to Chapter XXV, relative to 
some important air-lift work, recently done at a large mine in 
Mexico. The results of this work, which is still in progress as 
this edition goes to press, throw much light on the problems of 
raising water by air-lift under heavy heads, and show the 
adaptability of the method to unwatering deep mines. 

The author has availed himself of the opportunity to cor- 
rect such typographical errors as have been found in the previous 
issue of the book. 

R. P. 

June, 1920. 



PREFACE TO THE THIRD EDITION 



Since the third printing (2d Edition) of this book, in 
1913, there have been changes of some importance in the details 
of design of both compressors and rock-drills. Many of the 
illustrations have therefore been replaced in this edition by- 
cuts from current drawings furnished by the makers, and the 
necessary changes in the text have been made. 

In compressor design, the tendency during the last few years 
has been towards the larger use of " thin-plate " air valves, 
and an increase in the size and consequent efficiency of the 
intercoolers of stage compressors. For all except very small 
machines, stage compression is now the rule. Illustrations 
and descriptions of certain of the older compressors and drills 
have been retained in this edition, even though their makers 
consider them obsolete. This is thought advisable, because, 
while the builders may be giving especial attention to new or 
modified designs, large numbers of the older machines are still 
in use, and will continue to be employed for a long time to 
come. 

Among rock-drills, the old piston type has been little changed. 
Its field of work has become somewhat narrower, chiefly because 
of the improvements and larger use of hammer-drills for nearly 
all kinds of rock e:j;cavation. 

For this edition a number of the chapters have been entirely 
rewritten and several have been expanded. As it was not 
desired to increase the size of the book, added space has been 
made by condensation and omission of old matter. The 
matter contained in the appendix to the 1913 edition has 
been incorporated in Chapter XX. 

R. P. 

New York, July, 1918. 



PREFACE TO THE SECOND EDITION 



This edition has been revised and substantially enlarged. 
Among the principal additions are some 90 pages of text and 
63 illustrations, relating to the construction and operation of 
rock-drills, coal-cutting machines and channeling machines. This 
material is contained in Chapters XX, XXI, XXII, and XXIII. 
The detailed records of work of machine drills, in Chapters XX 
and XXI, I believe, will be found useful. Most of the data has 
not before been in print. 

In Chapter III the theory of the compression of air is pre- 
sented in greater detail, together with its applications to the 
operation and performance of compressors. The deductions of 
the more important formulae are also given, such as those used 
for calculating the horse-power required for single- and multiple- 
stage compression. In this connection I desire to acknowledge 
the kind assistance of Professor Charles E. Lucke, of Columbia 
University, and Professor H. J. Thorkelson, of the University of 
Wisconsin. To Dr. Lucke my thanks are due for the use of his 
valuable, and hitherto unpublished, notes relating to the work 
cycles of air compression, with and without clearance. I would 
call attention also to the records of compressor tests in the latter 
part of Chapter X. These comprise a few typical tests, selected 
from a large number recently made by Mr. R. L. Webb, Mechani- 
cal Engineer, on compressors of different kinds in a well-known 
Canadian mining district. 

Other new material has also been added, relative to the piston 
clearance of the air cylinders of compressors, and the ratio of 
inlet valve area to cylinder area. Numerous minor additions 
to the text have been made, together with corrections and alter- 

yii 



viii PREFACE TO THE SECOND EDITION 

ations where required. The new matter aggregates some 135 
pages of text and 87 illustrations. Many of the illustrations 
have been furnished by the respective makers of the machinery, 
to whom credit is duly given. In preparing this revision I have 
kept in mind certain kindly criticisms and suggestions received 
from readers of the first edition. R. P. 

New York, June, 1910, 



PREFACE TO THE FIRST EDITION 



The increasing use of compressed air makes the subject of 
interest to practitioners in nearly all branches of engineering. 
Besides its more important power applications, such as the 
operation of rock-drills, air brakes, riveting machines, and rail- 
road switching and signalling systems, the uses of compressed air 
are numerous in many minor branches of mechanical engineer- 
ing, in caisson work and the construction of subaqueous founda- 
tions, and in manufacturing industries, chemical works, etc., 
where it serves a multitude of purposes entirely distinct from that 
of the transmission of power. 

A realization of the breadth of the field has suggested that a 
book may be acceptable, addressed especially to those who are 
engaged in mining, tunnelling, quarrying, and other work involv- 
ing the excavation of rock, with its concomitant operations. 
While the literature bearing upon this branch of compressed-air 
service is by no means small, it is for the most part scattered 
through the technical periodicals and transactions of engineering 
societies, and therefore not readily accessible to those who are 
out of convenient reach of engineering libraries. I am aware that 
httle that is new can be said on this subject, and in writing the 
book I have availed myself freely of existing sources of informa- 
tion. 

In the first part, I have endeavored to present a view of cur- 
rent practice as to the construction and operation of compressors. 

Portions of the subject are dealt with at some length, for 
example : the types of compressor suitable for different kinds of 
service, heat losses occurring in air compression, and the various 
forms of valves, valve-motions, and governing and unloading 

135 



X PREFACE TO THE FIRST EDITION 

mechanisms, that constitute prominent features of modern com- 
pressor practice. A brief review is given of a few of the funda- 
mental principles of air compression, but my intention has been 
to present only enough of the theory to make intelligible the for- 
mulae employed for the ordinary calculations of the power and 
capacity of compressed air plant, together with the questions con- 
cerning temperature changes, as affecting the production and use 
of compressed air. Many details of the design of compressors and 
proportions of their parts have been omitted, since these fall 
properly within the province of the mechanical engineer. 

The second part is devoted to the applications (largely to 
mine service) of compressed-air transmission of power, including 
machine drills, pumps operated by compressed air, and mine 
haulage by compressed-air locomotives. 

Many of the illustrations are reduced or adapted from work- 
ing drawings kindly supplied by compressor builders. Others 
have been taken from catalogues of compressed-air machinery 
and from technical periodicals and books dealing with the 
different types. The origin of these has been stated in nearly 
every instance. My thanks are due to several of the technical 
journals, especially Compressed Air Magazine and Mines and 
Minerals, for many suggestions and in some cases for passages 
extracted either in substance or verbatim, from articles therein 
contained. For any important use or adaptation of pubhshed 
material, permission has been asked and obtained, and frequent 
references are given in foot-notes or in the body of the text. I 
also wish to acknowledge my indebtedness to Mr. Frank Richards' 
book on '' Compressed Air," from which I have derived sub- 
stantial assistance. 

Robert Peele. 

School of Mines, Columbia University, 
New York, May, 1908. 



CONTENTS 



PART FIRST 
PRODUCTION OF COMPRESSED AIR 

PAGE 

Preface iii 

List of Illustrations , . , » xiv 

CHAPTER I 

Introduction. Development of Air Compressors. Compressed Air versus 
Steam and Electric Transmission of Power i 

CHAPTER II 

Structure and Operation of Compressors: Straight-Line and Duplex; Com- 
pound Steam End; Stage Compressors; Direct- and Belt-Driven or 
Geared Compressors. Comparison of Types. Relation of Work Done 
in Air and Steam Cylinders. Proportions of Cylinders. Compressors 
Driven by Gasolene Engines, Water Wheels and Electric Motors 7 

CHAPTER III 

The Compression of Air. Outline of the Theory. Application of the Theory 
to the Operation of Air Compressors. Modes of Absorbing the Heat of 
Compression. Derivation of the Principal Formulas Relating to Iso- 
thermal and Adiabatic Compression 47 

CHAPTER IV 

Wet Compressors. Hydraulic-Plunger and Injection Compressors. Injection 
Apparatus. Quantity of Injection Water Required 72 

CHAPTER V 

Dry Compressors. Construction of the Water- Jackets. Circulation of Cool- 
ing Water. Piston Clearance and its Effect on Volumetric Capacity. 
Dry versus Wet Compression. Effect of Moisture in the Air under Com- 
pression. Effect of Injected Water 76 

xi 



xii CONTEXTS 

PaG£ 

CHAPTER VI 

Compound or Stage Compressors. Theor}' and Operation. Construction 
and Functions of the Intercooler. Deductions from the Indicator Card 
of the Stage Compressor 86 

CHAPTER \TI 

Air Inlet Valves. Chief Requisites of . Poppet Inlet Valves, their Construc- 
tion and Operation. Corliss Air Valves. IngersoU-Rand " Hurricane 
Inlet " Valve. Plate Valves. Le>Tier .Amiular Valve. Laidlaw-Dunn- 
Gordon '' Feather \'alve." IngersoU-Rand Plate Valve. Sullivan Plate 
\'alve. Arrangements for Admitting Inlet Air to the Compressor 99 

CHAPTER VIII 

Discharge or Delivery \'alves. Spring-Controlled Poppet Valves. Cataract- 
Controlled Poppets. Riedler Discharge Valve. " Thin Plate " Discharge 
Valves. Discharge Area for Air Cylinders 119 

CHAPTER IX 

Mechanically Controlled Air \'ahes and \'alve ^Motions. Considerations 
Regarding ^Mechanical Control for Inlet and Discharge Valves. Xorwalk, 
Xordberg, Laidlaw-Dunn-Gordon, Allis-Chalmers, Sullivan, Riedler, and 
Other Valve Motions. Koster Piston Valve 124 

CHAPTER X 

Performance of Air Compressors. Standards of Rating. Calculation of 
Horse-Power of Single-Cylinder and Stage Compressors. Mean CyUnder 
Pressure. Graphic Determination of Horse-Power. Temperature of 
Compression. Elements of Air Indicator Card. Compressor Tests. A 
Record of Field Tests, ^vith Diagrams and Tables of Horse-powers and 
Costs. Summary 137 

CHAPTER XI 

Air Receivers. Construction and Functions. Underground Receivers. Value 

of Cooling in the Receiver. Receiver ' After-Coolers " 169 

CHAPTER XII 

Speed and Pressure Regulators for Compressors. Speed Governors. Air 
Cylinder Unloaders. Modes of Regulation for Steam- and Belt-Driven 
Compressors. Constant-Speed, Variable Delivery Gear 175 

CHAPTER XIII 

Air Compression at Altitudes above Sea-Level. Consequent Reduction of 
Volumetric Capacity of Compressor. Relation between Compressor Out- 
put and Barometric Pressure. ]Mechanically Controlled Inlet Valves for 
High Altitudes. Stage Compression at High Altitudes 191 



CONTENTS xiii 

PAGE 

CHAPTER XIV 

Explosions in Compressors and Receivers. Discussion of Probable Causes. 
Heat of Compression. Cylinder Temperatures and Flashing-Points of 
Lubricating Oils. Examples of Explosions. Mode of Using Lubricant for 
Air Cylinders. Precautions for Preventing Explosions. Conclusions. . . 198 

CHAPTER XV 

Air Compression by the Direct Action of Falling Water. Principle and Theory'. 
Taylor Hydraulic Compressor. Descriptions of Plants at Magog, Prov- 
ince of Quebec, Ainsworth, B. C, Victoria Copper Mine, Mich., Cobalt, 
Ont., and Clausthal, Germany. Results of Tests 209 



PART SECOND 

TRANSMISSION AND USE OF COMPRESSED AIR 

CHAPTER XVI 

Conveyance of Compressed Air in Pipes. Loss of Power. Loss of Pressure. 
Discharge Capacity of Piping. D'Arcy's Formula. Graphic Solution 
of D'Arcy's Formula. Richards's Formula for Loss of Pressure. Com- 
parison of Results of Current Formulas. Compressed-Air Piping. Effect 
of Bends in Pipe-Lines 221 

CHAPTER XVII 

Compressed-Air Engines. General Considerations. Working at Full Pressure 
or with Partial or Complete Expansion. Ratios of Pressures and Tempera- 
tures due to Expansion in a Motor Cylinder. Corrections for Piston Clear- 
ance, etc. Nominal and Actual Cutoff. Cylinder Volume for a Given 
Power. Work Done in a Motor Cylinder. Volume of Free Air Required. 
Cummings' " Two-Pipe " System. Compressed-Air Hoists. Anaconda 
Mine Hoists. Other Examples 235 

CHAPTER XVIII 

Freezing of Moisture Deposited from Compressed Air. Causes and Preven- 
tion of Freezing. Influence upon Freezing of High Pressures in Trans- 
mission. Deposition of Moisture by Reduction of Pressure. Protection 
of Air Piping 253 

CHAPTER XIX 

Reheating Compressed Air. Appliances for, and Results of Reheating. Tem- 
peratures Employed and Consumption of Fuel. Construction and Opera- 
tion of Reheaters. Use of Reheaters for Underground Work. Wet and 
Pry Reheating 258 



XIV CONTENTS 

PAGE 

CHAPTER XX 

Compressed Air Rock-Drills. General Description. Drill Mountings. Clas- 
sification. Reciprocating Drills. Descriptions of " Sergeant," Wood, 
" Climax Imperial," Ingersoll-Rand " Butterfly," " Chicago Catling," 
Siskol, Murphy " Champion," Holman, Triumph, and Temple-Ingersoll 
" Electric-Air " Drills. Operation of Reciprocating Drills. General 
Considerations as to Efficiency of Machine Drills. Air Pressure. Con- 
sumption of Air, Normal and at Altitudes above Sea-Level. Factors 
Affecting Air Consumption, Examples from Practice. Proper Air Pres- 
sure for Machine Drills. Efficiency. Drill Repairs. Repair Costs. 
Records of Work. Conclusions 269 

CHAPTER XXI 

Hammer Drills. Classification. General Construction. Descriptions of 
Leyner-Ingersoll, Sullivan " Water " Drill, Stephen's " Climax," Hard- 
socg, Murphy, IngersollrRand " Jackhamers " and " Imperial," Sullivan 
* Rotators," [Chicago " Hummer " Drills, Waugh Drills, McKiernan- 
Terry Drills. Stope Drills of Different ]\Iakers. Air-Feed Drills. 
Operation of Hammer Drills. Air Consumption. Depth of Hole and 
Speed of Drilling. Records of Work. Field of Work. Makers of 
Hammer Drills 311 



CHAPTER XXII 

Coal-Cutting Machines. Classification. Endless Chain, Rotary-Bar, Disk 
and Pick Machines. Descriptions of Jeffrey and Sullivan Chain Cutters. 
Mavor and Coulson Disk Cutter. Pick Machines: Harrison, Sullivan 
and Ingersoll-Rand. Ingersoll-Rand Radia^axe," Sullivan " Post 
Puncher." Pneumelectric Pick Machine. Stanley Header. Auger Drills. 
Comparison of Coal Cutters. Loading Machines 350 



CHAPTER XXIII 

Channeling Machines. Applications and General Construction. Classifica- 
tion. Descriptions of Typical jMachines. Depth of Cut and Speed of 
Work. SL^es, Specifications, and Weights of Sullivan and Ingersoll- 
Rand Channelers. " Electric- Air " Channeler 380 



CHAPTER XXIV 

Operation of Mine Pumps by Compressed Air. Disadvantages of Using 
Ordinary Steam Pumps. Simple, Direct- Acting Pumps. Cylinder Dimen- 
sions of Simple Pumps. Volume of Air for Non-Expansive Working. 
Horse-Power. Regulation of Initial Air Pressure. Prevention of Freez- 
ing of ^loisture. Compressed-Air-Driven Compound Pumps. Discussion 
of Modes of Using the Air. Reheating for Compressed-Air Pumps 394 



CONTENTS XV 

PAGE 

CHAPTER XXV 

Pumping by the Direct Action of Compressed Air. Pneumatic-Displacement 
Pumps. Merrill and Latta-Martin Displacement Pumps. "Return- Air" 
System of Pumping. Air-Lift Pump, Theory and Operation. Volume 
of Air and Depth of Submergence. Design of Foot-Pieces. Application 
for Pumping Slimes. Lansell's Air-Lift for Pumping in Mine Shafts. 
Examples of Air-Lifts 409 

CHAPTER XXVI 

Compressed Air Haulage for Mines. Construction and Operation of Loco- 
motives. Modes of Dealing with Low Cylinder Temperature. Preheat- 
ing and Interheating. Specifications of Standard Locomotives. Con- 
struction Details. Calculations for Pipe-Line and Charging Stations. 
Design of Charging Stations. Calculation of Motive Power. Compres- 
sors for Charging Locomotives. Examples of Compressed Air Haulage 
Plants 432 

CHAPTER XX\TI 

Measurement of Air Consumption. Flow of Air through Orifices or Short 
Tubes. Measurement of Low-pressure Air. Measurement of High- 
pressure Air through Orifices, or by Meters or Tanks 467 



ILLUSTRATIONS 



PAGE 

Fig. I. — Laidlaw-Dunn-G ordon Straight-Line Compressor, i2"xi3"xi4" 

(Feather Valves) 9 

Figs. 2 and 3. — Ingersoll-Rand Straight-Line Compressor, Class " FR-i " 10, 11 

Fig. 4. — Sullivan Straight-Line Compressor, Class '' WA-5 " 12 

Fig. 5. — Sullivan Straight-Line, Two-Stage Compressor, with Simple Steam 

End, Class " WB-2 " Inset 12 

Fig. 6. — Norwalk Compound, Two-Stage, Straight-Line Compressor ... 13 
Figs. 7 and 8. — Sullivan Corliss Compound, Two-Stage, Straight-Line Com- 
pressor, Class " WC," 16" and 28"xi4|"and 24"x24" Cylinders , . . 14 
Figs. 9, 10 and 11. — Ingersoll-Rand Compound, Two-Stage Compressor, "Im- 
perial" Type, Class "XPV-3" 15,16,17 

Fig. 12. — Laidlaw-Dunn-Gordon Duplex, Cross-Compound, Two-Stage Com- 
pressor Inset 17 

Fig. 13. — Laidlaw-Dunn-Gordon Duplex, Cross-Compound, Two-Stage Com- 
pressor, with " Feather " Air Valves 18 

Fig. 14. — Auis-Chalmers Cross-Compound, Two-Stage Compressor (Corliss 

Inlet and Poppet Discharge Valves) 19 

Fig. 15. — Combined Air and Steam Cards, Straight-Line Compressor ... 23 
Fig. 16. — ^IngersoU-Rand Portable, Gasolene-Driven Compressor .... 26 
Fig. 17. — Straight-Line, Gasolene-Driven Compressor (Chicago Pneumatic 

Tool Co.) 27 

Fig. 18. — Straight-Line, Gasolene-Driven Compressor, Class "ED-i" (Inger- 
soll-Rand Co.) 28 

Fig. 19. — Duplex Compressor, with i6"x3o" Cylinders, Direct-Connected to a 
i6-ft. Risdon Waterwheel, Working under 300-ft. Head. Built by Risdon 

Iron Works Co. for the Goleta Mining Co 29 

Fig. 20. — Nordberg Compressor, Direct-Connected to a Pelton Wheel ... 30 

Fig. 21. — Pelton Impulse Wheel . . 31 

Fig. 22. — Needle Nozzle for Pelton Wheel ' -34 

Fig. 23. — Sullivan Straight-Line, Two-Stage, Belt-Driven Compressor, Class 

*'WH-3," 12" and 7f' xio" Cylinders 35 

Fig. 24. — Laidlaw-Dunn-Gordon Straight-Line, Two-Stage, Belt-Driven 

Compressor, 14" and 8J"x 1 2" Cylinders, Feather Air Valves .... 36 
Fig. 25. — Sullivan Belt-Driven, Two-Stage, Angle Compressor, Class " WJ-3 " 

(1917) .- • ■ 2>7 

Fig. 26. — Alley & McClellan Vertical, Two-Stage, Belt-Driven Compressor 

(from paper by G. Blake Walker, Trans. Midland Inst. Min., Civ., and 

Mech. Engs., Jan. 21, 1913) 38 

Fig. 27. — Ingersoll-Rand Duplex, Two-Stage, Belt-Driven Compressor, with 

Plate Valves (high-pressure side shown) 39 

xvii 



Xviii ILLUSTRATIONS 

PAGE 

Fig. 28. — Ingersoll-Rand Compressor, Chain-Driven from an Electric Motor, 

Class "NE-i," 1 2i"x 1 2" Cylinder 40 

Fig. 29. — Ingersoll-Rand Direct-Connected, Motor-Driven, "Imperiia.]" 
Duplex Compressor, Type 10 (longitudinal section through low-pressure 

cylinder) Inset 40 

Fig. 30. — Robey & Co. High-Speed, Vertical, Two-Stage, Motor-Driven 

Compressor 41 

Fig. 31. — Turbo Compressor, Driven by a Westinghouse Steam Turbine, 

New Hucknall Colliery, Nottinghamshire, England (G. Blake Walker) . 43 
Fig. 32. — Series of Impellers, Westinghouse Turbo Compressor .... 44 
Fig. ;^^. — British- Westinghouse Rateau Mixed-Pressure Turbine and Air- 
Compressor. Speed, 4,000 r.p.m., output 7,500 cu. ft. free air per min. 
at 80 lbs. gauge. Tests by W. F. Mylan {Trans. Instn. Min. Engs., 

England, Vol. 45, pp. 245-264) 45 

Fig. 34. — Compression-Temperature Diagram 52 

Fig. 35. — Reference Diagram, Elements of Air-Indicator Card • • • • 55 

Fig. 36. — Air-Indicator Card 58 

Fig. 37. — Reference Diagram, Two-Stage Compressor, No Clearance ... 61 

Fig. 38. — Single-Stage Compressor Diagram, with Clearance 63 

Fig. 39. — Two-Stage Diagram, with Proportionate Clearance 66 

Fig. 40. — Two-Stage Diagram, with Disproportionate Clearance .... 68 

Fig. 41. — Humboldt Wet Compressor 73 

Fig. 42. — Air Cylinder of NOrdberg Compressor 77 

Fig. 43. — Air Cylinder of Allis-Chalmers Compressor, with 9 Delivery Valves 

Set Radially at Each End 78 

Fig. 44. — Air Cylinder, Class E Compressor, Laidlaw-Dunn-Gordon Co. . . 79 

Fig. 45. — Air Card Showing Effect of Piston Clearance 80 

Fig. 46. — Diagram Showing Effect of Piston Clearance {Eng. News) ... 81 
Fig. 47. — Section of Air Cylinder, Showing Method of Reducing Clearance . 82 

Fig. 48. — Diagram of Norwalk Two-Stage Compressor 89 

Fig. 49. — Ingersoll-Rand Intercooler, with Steel Shell and Cast-iron Water 

Heads 94 

Fig. 50. — Vertical Intercooler. Ingersoll-Rand Co 96 

Fig. 51. — Combined Air Cards of Two-Stage Compressor 97 

Fig. 52. — Norwalk Poppet Inlet Valve 102 

Fig. 53. — Laidlaw-Dunn-Gordon Poppet Inlet 102 

Fig. 54. — Diagram Showing Effect of Valve-Spring Resistance on Volumetric 

Capacity of Compressors {Eng. News) 104 

Fig. 55. — Air Card Showing Effect of Valve Resistance 105 

Fig. 56. — "Skip- Valve." Norwalk Iron Works Co 106 

Fig. 57. — Cylinder of " Hurricane-Inlet " Compressor. Ingersoll-Rand Co. . 108 

Fig. 58. — "Hurricane-Inlet" Valves. Enlarged Section 109 

Fig. 59. — Leyner Compressor. Part Section, Showing Fiat Annular Air 

Valves Ill 

Fig. 60. — Leyner Annular Inlet Valve 112 

Fig. 61. — Laidlaw-Dunn-Gordon Air Cylinder, with Valves in Place .112 

Fig. 62. — Laidlaw-Dunn-Gordon "Feather" Valve, with Cover Removed . 113 



ILLUSTRATIONS xix 

PAGE 

Fig. 63. — Laidlaw-Dunn-Gordon "Feather" Valve, Assembled . , . .113 
Fig. 64. — Cylinder and Intercooler of Ingersoll-Rand Class "ORG" Com- 
pressor, with Plate Valves 115 

Fig. 65. — Parts of the Ingersoll-Rand Plate Valve 115 

Fig. 66.^-Section of Ingersoll-Rand Plate Valve, Assembled 115 

Fig. 67. — Sullivan Plate Valve, Low-Pressure Cylinder, Angle-Compound 

Compressor, Class " WJ-3 " .116 

Fig. 68. — Sullivan Plate Valves in Place on Inner Head of Low-Pressure 

Cylinder, Angle- Compound Compressor, Class "WJ-3" 117 

Fig. 69. — Ingersoll-Rand " Imperial " Poppet Discharge Valve . , . . .120 

Fig. 70. — Laidlaw-Dunn-Gordon Poppet Discharge Valve 120 

Fig. 71. — Norwalk Poppet Discharge VaWe 121 

Fig. 72. — "Express" Poppet Valve. Riedler Compressor 122 

Fig. 73. — Valve Motion of Low-Pressure Air Cylinder, Norwalk Compressor 1 26 

Fig. 74. — Air Cylinder of Nordberg Stage Compressor 128 

Fig. 75. — "Cincinnati" Valve Gear. Laidlaw-Dunn-Gordon Compressor . 129 

Fig. 76. — Standard Air- Valve Motion. AUis-Chalmers Co 130 

Fig. 77.— Sullivan Air Cylinder, Corliss Inlet and Poppet Discharge Val^^es . 131 

Figs. 78, 79 and 80. — Riedler Air-Valves 133, 134, 135 

Figs. 81 and 8ia.— Nomograms for Graphic Determination of Compressor 

Horse-Power, Single- and Two-Stage 142, 143 

Fig. 82. — Diagram. Elements of Air-Indicator Card ....... 146 

Fig. 82a. — Air Card Diagram 147 

Fig. 83. — Combined Cards, Two-Stage Nordberg Compressor 151 

Fig. 84. — Combined Cards, "Imperial, Type 10" Two-Stage Compressor . 152 
Fig. 85. — Card from 30j"x24" Low-Pressure Air Cylinder of Style "O," 

Ingersoll-Rand Compressor 153 

Fig. 86.— Card from i8|"x24" High-Pressure, Air Cylinder of Style "O," 

Ingersoll-Rand Compressor 153 

Figs. 87, 88 and 89. — Curve Diagrams, Compressor Plant No. i . 154, 155, 157 
Figs. 90, 91 and 92. — Curve Diagrams, Compressor Plant No. 2 160, 161, 162 
Figs. 93, 94 and 95. — Curv^e Diagrams, Compressor Plant No. 3 . 165, 166, 167 

Fig. 96. — Curve Diagram, Compressor Plant No. 4 168 

Fig. 97.— Vertical Air Receiver, Norwalk Iron Works Co 169 

Fig. 98. — Horizontal Receiver- Aftercooler, Ingersoll-Rand Co 170 

Fig. 99. — Ingersoll-Rand Vertical Aftercooler, Type "VK" 173 

Fig. 100. — Clayton Governor and Pressure Regulator 176 

Fig. ioi. — Norwalk Pressure Regulator 177 

Fig. 102. — Sullivan Pressure Regulator . . . 179 

Fig. 103. — Laidlaw-Dunn-Gordon Air Governor i8i 

Figs. 104 and 105. — Ingersoll-Rand "XPV" Automatic Air Governor 182, 183 

Fig. 106. — Rand Imperial Unloader 184 

Figs. 107 and 108. — Nordberg Constant-Speed, Variable-Delivery Compressor, 

Valve-Motion and Regulating Gear 186, 187 

Fi6s. 109, 1 10 and in . — Indicator Cards, Nordberg Constant-Speed, Variable- 
Delivery Compressor 189 

Fig. 1X2.— Ingersoll-Rand "RA-39" Controller 190 



XX ILLUSTRATIONS 

PAGE 

Fig. 113. — Air Cards Showing Results of Compression at Altitudes above Sea- 
Level 192 

Fig. 114.— Taylor Hydraulic Air Compressor 211 

Fig. 115. — Taylor Hydraulic Air Compressor, Detail of Head-Piece .212 

Figs. 116 and 117. — Hj^draulic Air-Compressing Plant at Kootenay, British 

Columbia Inset 213 

Fig. 118. — Hydraulic Air Compressing Plant. Victoria Mine, Mich. Inset 216 
Fig. 119. — Cobalt Power Co.'s Plant, Plan and Elevation .... Inset 217 
Fig. 120. — Hydraulic Air Compressor at Clausthal, Germany . . . Inset 219 
Figs. 121 and 122. — Diagrams for the Graphic Solution of D'Arcy's For- 
mula 228, 229 

Fig. 123. — Expansion Curves of Steam and Air 237 

Fig. 124. — Card Showing Work Done in Motor Cylinder 241 

Fig. 125.— Ingersoll-Rand "Little Tugger" Hoist, Type "i-H" . . . .249 

Fig. 126. — Ingersoll-Rand Reheater 263 

Fig. 127. — Old Rand Reheater 265 

Fig. 128. — Sullivan Reheater 265 

Fig. i29.^Tripod Mounting, Sullivan Drill 271 

Fig. 130. — Double-Screw Column Mounting 272 

Fig. 131. — "Sergeant" Rock Drill, Ingersoll-Rand Co 275 

Fig. 132. — Spool- Valve and Chest, "Sergeant" Rock Drill 276 

Fig. 133, — " Sergeant " Ratchet and Rifle-Bar 277 

Fig. 134. — "Liteweight" Drill, with Water Attachment. Sullivan Machinery 

Co. 277 

Fig. 135. — Rotation Device, Sullivan Drill 278 

Fig. 136.— Sullivan Drill Mounted, with Water Tank and Connections. . . 279 

Fig. 137.— "Wood"' Piston Drill. Wood Drill Works 280 

Fig. 138. — Stephens "Climax Imperial" Drill, 3^". R. Stephens & Son, 

Cam Brea, Cornwall, England 281 

Fig. 139. — Holman Spool- Valve Drill 283 

Fig. 140. — Ingersoll-Rand "Butterfly" Drill, for Tripod or Column Mounting 284 

Fig. 141.— Valve of "Butterfly" Drill 284 

Figs. 142 and 143. — Diagrams of Operation of " Butterfly " Drill .... 285 

Fig. 144. — " Chicago Catling " Drill. Chicago Pneumatic Tool Co . . 286 

Fig. 145.— Siskol Drill (2^^") 287 

Fig. 146.— Sullivan "Hy-Speed Drill, with Water Attachment 288 

Fig. 147. — Murphy "Little Champion" Drill (Wickes Machinery Co.) . . 289 

Fig. 148.— Temple-Ingersoll "Electric-Air" Drill 291 

Fig. 149.— "Triumph" Drill 294 

Figs. 150 and 151. — Leyner-Ingersoll Water Drills, Nos. 18 and 26 . . . 314 

Fig. 152. — Rotation Device of Leyner-Ingersoll Drill, No. 18 315 

Fig. 153. — Details of Bit Shanks 31.S 

Fig. 153a. — Leyner-Ingersoll Water Drill, Nos. 148 and 248 317 

Fig. 154.— Sullivan "DR-6" Drill, with Water Attachment 317 

Fig. 155. — Stephens " Climax Imperial " Hammer Drill Inset 318 

Fig. 156. — Hardsocg Wonder Drill, with D-Handle 320 

Fig. 157. — Hardsocg Drill, with Automatic Rotation 321 



ILLUSTRATIONS xxi 

PAGE 

Fig. 158.— IngersoU-Rand ''Jackhamer" Hand Drill, "BCR-430," for Dry 

Holes 323 

Fig. 159. — Rotation Device of " Jackhamer " Drill 323 

Fig. 160. — IngersoU-Rand "Water Jackhamer," for Wet Holes .... 323 

Fig. 161. — "Jackhamer" Drill on Cradle Mounting 324 

Fig. 162. — "Jackhamer" on Special Mounting for Thin Coal Seams . . . 324 

Fig. 163. — Twisted Cruciform Bit, foi Drilling in Coal , 325 

Fig. 164. — "Jackhamer Sinker" 326 

Figs. 165 and 166. — " BuUmoose Jackhamer " 326, 327 

Fig. 167. — IngersoU-Rand "Imperial" Hammer Drill, Types "MV-i" and 

"MV-2" 328 

Fig. 168.— Sullivan "Air-Tube Rotator" ("DP-33") 329 

Fig. 169.— Sullivan "Water-Tube Rotator" ("DP-33") 330 

Fig. 170.— Sullivan "Auger Rotator" Class "DR 33" 331 

Figs. 171 and 172. — Mountings for Sullivan " Rotators " .... 332,333 

Fig. 173. — "Hummer" Drill. Chicago Pneumatic Tool Co 334 

Fig. 174. — Valve Mechanism of " Hummer " DriU (diagrammatic) . . . 335 

Fig. 175.— Waugh "Clipper" Drill, Model so 336 

Fig. 176. — McKiernan-Terry " F-i " Hammer DriU 338 

Fig. 177. — McKiernan-Terry "A-9" Hammer DriU . . 339 

Fig. 178.— Waugh Stoper, Model "14-A," Denver Rock DrUl Mfg. Co. . 341 
Figs. 179 and 180. — Waugh Stoper. Diagrammatic Sections of Cylinder, 

Valve and Ports 342 

Fig. 181. — IngersoU-Rand "Butterfly" Stoper 343 

Fig. 182. — Oiling Device, "Butterfly" Stoper 344 

Fig. 183.— Air Feed for "Butterfly "Stoper, Type" BC-2 1 " . . . . .345 
Fig. 184. — IngersoU-Rand "CC Stopehamer" (Dry Type) ...... 345 

Fig. 185. — Cylinder and Valve of "CC Stopehamer," 345 

.Fig. 186. — Dust Allayer for " Stopehamer " 346 

Figs. 187 and 188. — Modes of Operating Sullivan Coal Cutters in Room and 

Longwall Work 351^352 

Fig. 189. — Jeffrey Coal Cutter. Methods of Manipulation in Starting a Room 

Cut 353 

Fig. 190. — Longwall Air-Turbine Coal Cutter, Type " 24-A," Jeffrey Mfg. Co. 354 

Fig. 191. — Compressed- Air Turbine, Jeffrey Mfg. Co 355 

Fig. 192. — Diagram of Operating Mechanism, Jeffrey Coal Cutter ("24-A") 355 

Fig. 193. — Jeffrey Shortwall Coal Cutter, Type "3 5-B" 356 

Fig. 194. — Jeffrey Chain Machine, Type " 1 6-D" 358 

Fig. 195. — Sullivan " Ironclad " Longwall Coal Cutter, Class " CH-8 " . . 359 
Fig. 195a. — Sullivan Room-and-Pillar Chain Machine, Class "CE-7" . . 360 

Fig. 196. — Rotary-Bar Coal Cutter, Mavor & Coulson 362 

Fig. 197. — Mode of Operating the Mavor & Coulson Bar Cutter . . . . 363 

Fig. 198. — ^Jeffrey Disk Coal Cutter, Style "2 2-C" 364 

Fig. 199. — Sullivan Coal Pick, Working in a Thin Vein 365 

Fig. 200.— Harrison Pick Machine, Types "PG" and "PW" 367 

Fig. 201. — Sullivan Coal Pick 368 

Fig. 202. — IngersoU-Rand Coal Pick . ., 369 



xxii ILLUSTRATIONS 

PAGE 

Fig. 203. — Ingersoll-Rand Coal Pick, Diagram of Valves and Ports . . . 370 

Fig. 204. — Ingersoll-Rand " Radialaxe " Coal Pick 371 

Fig. 205. — Sullivan "Post Puncher" 372 

Figs. 206 and 207. — Pneumelectric Coal Puncher 373 

Fig. 208. — Diagram of Gearing, Pneumelectric Coal Puncher 374 

Fig. 209. — Stanley Heading Machine for Collieries 375 

Fig. 210. — Jeffrey Air Loader for Collieries, Class ''38-A" 378 

Fig. 211. — Sullivan Track Channeler 381 

Fig. 2X2. — Ingersoll-Rand Ram Track Channeler, for Marble 382 

Fig. 213. — Sullivan Rigid Back, Steam-Driven Channiler 383 

Fig. 214.— Sullivan Adjustable Back, Air-Driven Channeler 385 

Fig. 215. — Ingersoll-Rand Undercutting Track Channeler, Type "HF-3" . 386 

Fig. 216. — Ingersoll-Rand "Broncho" Channeler 387 

Fig. 217. — Gibson-Ingersoll "Electric-Air" Track Channeler, with Swing Back 392 

Fig. 218. — Merrill Pneumatic Displacement Pump 410 

Fig. 219. — Latta-Martin Pneumatic Displacement Pump 412 

Fig. 220. — "Return- Air" Displacement System, Ingersoll-Rand Co . . . 414 

Fig. 221. — Diagram of Pohle Air-Lift Pump 416 

Fig. 222. — Foot-Pieces and Modes of Piping Air-Lift Pumps 423 

Fig. 223.— Foot-Pieces, Types "VA" and "VC," Ingersoll-Rand Co . . . 424 

Fig. 224. — Foot-Piece for Air-Lift Pump, for Raising Mill Tailings and Slimes 425 

Fig. 225. — Diagram of Lansell's Air-Lift Pump for Mine Shafts .... 428 

Fig. 226. — Air-Lift at Old Dominion Copper Mine, Globe, Arizona . . . . 429 

Fig. 227. — Two-Stage Air-Lift, Coolgardie, W. Australia 430 

Fig. 228.— H. K. Porter Co.'s Locomotive, Class "B-P-0" 435 

Fig. 229. — Diagram of Reheater and Cyhnders for H. K. Porter Co.'s Two- 
Stage Locomotives (Table L) 436 

Fig. 230. — Baldwin Four-Wheel Compressed- Air Locomotive , . , . 437 

Fig. 231. — Baldwin Six-Wheel Compressed- Air Locomotive 437 

Fig. 232.— H. K. Porter Co.'s Locomotive, Class "B-PP-0" 438 

Fig. 233.— H. K. Porter Co.'s Locomotive, Class "C-PP" 439 

Fig. 234. — H. K. Porter Co.'s Locomotive for Small-Scale Work .... 443 
Fig. 235. — Diagram of Piping for Interheater, Valves, etc., on Rear End of 

Porter Two-Stage Locomotives 444 

Fig. 236. — Baldwin Single-Stage Locomotive, 9 xi4-in. Cylinders . . Inset 444 

Fig. 237. — H. K. Porter Co.'s Locomotive, Class " C-5PS-O " 445 

Fig. 238.— H. K. Porter Co.'s Locomotive, Class "B-PP-T" 446 

Fig. 239. — Compressed- Air Locomotive Charging Station 450 

Fig. 240. — Charging Station, H. K. Porter Co 451 

Fig. 241. — Norv.'alk Locomotive-Charging Compressor 454 

Fig. 242. — Air-End of Ingersoll-Rand Three-Stage Locomotive Charger . . 455 
Figs. 243 and 244. — Low- and High-Pressure Air-Ends of Ingersoll-Rand Four- 
Stage Compressor 456 

Fig. 245. — Ingersoll-Rand Four-Stage Compressor 457 

Fig. 246. — E. A. Rix Compressed-Air Locomotive for Empire Mine, Grass 

Valley, California 460 

Fig. 247. — Diagram of Air Apparatus for Measuring Low-pressure Air. . .471 



ILLUSTRATIONS XXlll 

PAGE 

Fig. 248. — Manifold and Nozzles for Measuring the Discharge of Compressed 

Air 473 

Fig. 249. — The "Tool-om-eter," for Measuring Compressed Air .... 477 

Fig. 250. — Measuring Tanks for Rock-Drill Testing . 479 

Fig. 251.— Tanks for Measuring Air Consumption of Rock-Drills . . . 481 
Fig. 252. — Air Receiver .\rranged for Measuring Air from a Compressor . 481 

Fig. 253. — Time-Pressure Curves 482 

Fig. 254. — Diagram of Tank for Measuring Air Consumption of Rock-Drills . 48 



COMPRESSED AIR PLANT 



Part First 
PRODUCTION OF COMPRESSED AIR 



CHAPTER I 
INTRODUCTION 

One of the most important applications of the transmission of 
power by compressed air is the driving of machine rock-drihs; 
and to the necessity of providing for these drills a power medium 
suitable for use in mines and tunnels has been due, more than to 
any other cause, the development of the modern air compressor. 

The time which has elapsed since the beginnings of this branch 
of engineering is short. The first percussion rock-drill, operating 
independently of gravity, was invented in 1849 by J. J. Couch, of 
Philadelphia. Though used only experimentally, it embodied 
the principal mechanical features of the modern machine-drills, 
which have had such a striking influence in mining and tunnel- 
ling. Couch's machine, together with its immediate successors, 
such as the Fowle drill (1849-51) and the Cave (Paris, 185 1), 
were steam-driven and therefore unsuitable for underground 
work. In 1852, the physicist Colladon proposed the use of com- 
pressed air for operating rock-drills, in connection with the 
driving of the Mont Cenis tunnel, in the western Alps. His idea 
was developed by Sommeiller and others between 1852 and i860, 
?Lnd in 1861-62 an air-compressor plant was first used successfully 



2 COMPRESSED AIR PL.\NT 

at that tunnel. It was driven by water power and furnished air 
for ventilation as well as for the drills. The early air compres- 
sors were crude in design. Sommeiller's first plant, though of 
large size, had some resemblance in principle to the old hydrauhc 
ram, possessing no moving parts except the valves. Steam- 
driven piston compressors, as the Dubois-Frangois, and more 
or less similar to some of the wet compressors still in use in 
Europe, soon made their appearance. The first compressors 
built in the United States were the Burleigh, employed at the 
Hoosac Tunnel, on the Boston and Albany Railroad, in 1865-66. 
The Norwalk, Clayton, and Rand compressors were among the 
earlier makes in this country. 

In Europe, the Mont Cenis tunnel, about eight miles long 
and completed in 187 1, the first connecting link through the 
Alps between the railway systems of France and Italy, v/as the 
field where were solved on a large scale many problems of 
compressed-air production and use. Sommeiller there laid the 
foundations of new practice, by which that great work was 
successfully completed. From 1857 to 1861 the tunnel headings 
had been progressing slowly and in the face of great difficulties. 
Drilling was done by hand labor and blasting by black powder, 
the average advance for this period, in each of the two headings, 
being only about 1.5 ft. per day. At this rate, granting that 
the work could have been finished at all by the means employed, 
over 40 years would have been required to connect the headings 
and years more to complete the enlargement to full section. 
With machine drills, the average speed of advance in each 
heading rose to 4.75 ft. per 24 hrs. and later, when dynamite 
was introduced, to a little over 6 ft.; this average being main- 
tained for a period of 6 years. 

Machine drills did not make their way into mining to any 
extent for some }'ears after their successful application to tunnel- 
driving. It is difficult now to name the mining district in this 
country where they were first used, but their most important 
trial was in the Calumet and Hecla copper mine, Michigan. 
After strong opposition from the miners, the Rand drill was 
introduced there in 1878, and the value of machine driUing was 



INTRODUCTION 3 

soon demonstrated by decreased costs of drifting and stoping 
and higher speeds of advance. 

Compressed air has now a wide appHcation in various branches 
of mechanical engineering. In this book it is intended to deal 
only with its production and uses for mining and tunnelling 
operations. Its two rivals in these fields of work are steam 
and electricity. 

As compared with steam, compressed-air transmission of 
power is useful and convenient for three reasons: first, its loss 
in transmission through pipes is relatively small; second, the 
troublesome question of the disposal of exhaust steam under- 
ground is avoided; third, the exhausted air is of some assistance 
in ventilating the working places of the mine. In large mines, 
where steam may be carried thousands of feet, down shafts and 
through lateral workings, for operating pumping engines, etc., 
the disadvantages attending its use are apparent; condensation 
is serious, even when the piping has good non-conducting cover- 
ing, and the efficiency becomes abnormally small. Furthermore, 
aside from the heat produced by the use of steam, it is rarely 
feasible to employ efficient condensers for underground engines 
other than pumps, because of the difficulty of obtaining con- 
densing water. If the exhaust be discharged into the mine 
workings, even though these are large and well ventilated and 
the volume of the exhaust steam comparatively small, the 
temperature and quantity of moisture in the air is considerably 
increased. Deterioration of the timbering is hastened, the roof 
and walls of the workings are often softened and slacked off, 
and the mine atmosphere is rendered oppressive and unwhole- 
some. The presence of hot steam pipes in confined workings, 
or in the narrow compartments of shafts, is also objectionable. 

Even with use of the best non-conducting covering the 
condensation loss in long steam lines greatly reduces the effective 
pressure at a distant underground engine and very uneconomi- 
cal working is the result. In conveying steam several thousand 
feet the pressure may be reduced to half the boiler pressure, 
or less. Thus, in the case of a pump situated 2,000 ft. from 
the boiler and using 200 cu.ft. of steam per minute at a boiler 



4 COMPRESSED AIR PLANT 

pressure of 75 lbs., with a 4-in. mineral- wool-covered pipe, the 
effective pressure at the pump would be only about 58 lbs.; 
or, with a poor covering, like some of the asbestos lagging often 
used, it might be as low as 35 lbs. In compressed-air trans- 
mission, on the other hand, the reduction of pressure for the same 
volume of air, size of pipe, and initial pressure, would be 9.3 
lbs., giving a terminal pressure of 65.7 lbs. However, as the 
speed of flow in pipes for economical transmission is greater for 
steam than for air, a comparison based solely on piping of the 
same diameter cannot justly be made. In the above example, 
if the diameter of the pipe were smaller the gain in reduced 
radiation would outweigh the increased frictional loss, and the 
net loss would be diminished. Since the frictional loss varies 
inversely, and the loss from radiation directly, with the diameter, 
the size of the steam pipe can be so proportioned as to produce 
a minimum loss under given conditions. With compressed air 
the case is different, since the question of radiation is eliminated. 
If the pipe diameter be increased to 5 ins. the loss of pressure, 
or head required to overcome friction, is reduced to 2.8 lbs., and 
increasing the distance to one mile it would be only 7.4 lbs. 
Furthermore, as against the increased cost of the larger air pipe, 
there is the expense of the non-conducting covering necessary 
for steam transmission. 

Thus, compressed air may be conveyed long distances with 
but small loss of pressure; it is always ready to do its work, 
and, aside from leakage of pipes, which is preventable, it suffers 
no loss of power when not in actual use. For performing work 
intermittently, at a distance from its source, it is therefore particu- 
larly valuable, because the air pressure is maintained nearly 
constant during intervals of work, without further expenditure 
of power. With steam transmission, power is continually dis- 
sipated by radiation, and a steam engine, when stopped for any 
length of time, loses much of its normal working temperature 
and becomes a receptacle for water of condensation. 

Though compressed air is employed in mining mainly for 
operating machine drills, it is used also for underground hoists 
and pumps, and sometimes for mechanical coal cutters, in both 



INTRODUCTION 5 

bituminous and anthracite mines. Compressed-air locomotives 
in mines and tunnels exemplify its capacity for storing power, 
in contradistinction to its function as a power transmitter. 
The introduction of machine drills has facilitated the driving of 
railroad and mine tunnels, which otherwise would have been 
greatly delayed or completed only with difficulty. Had com- 
pressed-air power, together' with the high explosives, not been 
available, it is doubtful whether the great tunnels in the Alps 
and elsewhere^ and the numerous long mine tunnels driven in 
recent years in this country, would have been at all practicable. 

Without reviewing in detail the comparative merits of 
electricity and compressed air, it may be pointed out that the 
application of electricity for transmitting power in mines has 
increased enormously in importance since about 1888. The 
peculiar requirements of mine service have been in most cases 
successfully met by modifications of standard forms of electric 
apparatus. Both means of power transmission possess char- 
acteristics which adapt them particularly for underground work. 
But, although electricity rivals compressed air in nearly all 
branches of mine work, their spheres of usefulness are not identical 
and the field is broad enough for both. Though it is sometimes 
stated that the first cost of an electric plant is lower than that of 
an equivalent compressed-air plant, there is actually but little 
difference between the costs of the plants themselves. For 
short-distance transmission of a given power an electric con- 
ductor line costs much less than compressed-air piping; but 
the cost of the electric line increases as the square of the distance, 
while that of the pipe line increases only as the first power of 
the distance. It is in the greater efficiency of generation that 
electric transmission has the most marked advantage. 

In one direction only has electricity failed hitherto to meet 
every requirement. While compressed-air drills, though far 
from being economical machines, nevertheless admirably fulfil 
their purpose, no satisfactory electric rock-drill has yet been 
produced. It is to be hoped that a solution of this problem 
may be found. The Temple-Ingersoll " electric-air " drill, 
brought out about 1902, is an ingenious machine, but not an 



6 COMPRESSED AIR PLANT 

electric drill in the ordinary meaning of the term. It is a com- 
bination of a compressed-air drill of special design, operated by 
a small, electric-driven compressor. As there is no exhaust, 
an incidental advantage of the ordinary air drill is missing, 
namely, that of assisting somewhat in ventilating those places 
where ventilation is most needed. This, together with such 
minor uses of compressed air as the cleaning of drill holes pre- 
paratory to charging, and driving out the smoke of blasting 
from working places, renders it doubtful whether, for under- 
ground mining, electric drills of any kind can supersede entirely 
those operated by compressed air. Given the necessity for a 
compressed-air plant for rock-drills, as is the case in most 
metal mines, it may often be more advantageous to provide the 
additional compressor capacity required for driving underground 
pumps, hoists, and other machines as well, than to install a 
separate plant for generating electricity. 

It has long been customary to consider compressed air 
as a mode of power transmission respecting which the questions 
of convenience and expediency are more weighty than the 
attainment of a high degree of efficiency. But, as the principles 
of air compression have become better understood, a substantial 
improvement has taken place in compressor design, the instal- 
lation of pipe lines, and the operation of machines using com- 
pressed air. The consequences of overloading a compressor, 
and thereby driving it beyond its proper speed, are comprehended 
by every intelligent master mechanic as being wholly different 
from those due to overloading a steam engine. The results 
of leaks in air pipes, and of using air mains of too small a diam- 
eter, are also understood. By better practice in the production, 
transmission, and use of compressed air a higher total efficiency 
is now realized than formerly was thought possible. 



CHAPTER II 
STRUCTURE AND OPERATION OF COMPRESSORS 

An air compressor consists of a cylinder in which atmospheric 
air is compressed by a piston, the driving power being derived 
from a steam engine, water-wheel, or electric motor. The air- 
cyhnder is usually double-acting, with inlet and delivery valves 
in each head. The air is compressed by the advancing piston, 
while, in the simplest compressors, the decrease in pressure, or, 
as it is commonly termed, the tendency to a vacuum, behind 
the piston causes the inlet valves to open under atmospheric 
pressure, thus allowing air to flow into the cylinder. At each 
stroke a certain volume of compressed air is forced out through 
the discharge valves, into a pipe leading to a reservoir or receiver^ 
whence the air enters the transmission pipe or main. 

No single classification of air compressors can be made 
sufficiently comprehensive to present all of their salient features. 
Three bases of comparison suggest themselves. First, the 
structural characteristics of compressors regarded purely as 
engines; second, the mode of dealing with the heat produced 
during compression; third, the numerous types of air valves 
and valve-motions. The first classification is given here, the 
others being taken up respectively in Chap. IV-VI and VII-IX. 
Air-brake and gas compressors, vacuum pumps and other special 
air-compressing machinery are omitted, as this book deals only 
with compressors which are applicable to mine or similar service. 
(A classification including compressors for nearly all kinds of 
service, as built by an important American maker, is given at 
the end of this chapter.) 

First Classification, taking the steam-driven compressor 
as the type form: 

7 



8 COMPRESSED AIR PLANT 

1. ** Straight-line '* Compressors. The steam and air cyUn- 
ders are set tandem on a common piston-rod. There is a pair 
of fly-wheels, one on each end of the crank-shaft, driven by 
a single connecting rod, or, in some designs, by outside con- 
necting-rods from a cross-head between the cylinders. (Figs. 
I to 4.) 

2. Duplex Compressors. Two engines are placed side by 
side, each consisting of tandem steam and air cylinders, with 
their cranks set at 90° on a common fly-wheel shaft. Each side 
of the duplex is in effect a straight-line compressor. The steam 
cyhnders may be simple or compound; the air cylinders single 
or staged. (Figs. 9-14.) 

3. Compressors with Compound Steam Ends, (a) Duplex, 
horizontal, cross-compound; a single-stage air cylinder being 
set tandem to each steam cylinder. This form is now rare. 
The considerations leading to the compounding of the steam 
end make it desirable to use stage compression. Nearly all 
large, steam-driven compressors are now of the duplex, cross- 
compound, two-stage type (Figs. 9-14). (b) Vertical compound; 
the air cylinders being placed above the steam cylinders. This 
also is an unusual design. Some large King-Riedler com- 
pressors,* up to a capacity of 8,000 cu.ft. of free air per minute, 
have been built for South African mines. In Great Britain 
vertical compressors, both large and small, have had considerable 
vogue in recent years. Prominent among them are those of 
Belliss & Morcom, Peter Brotherhood Co., Alley & MacClen- 
nan, and Robey & Co. 

The chief advantage of vertical compressors is the saving of 
floor space, which is rarely of consequence at mines. Disad- 
vantages are the relative inaccessibility of the working parts, 
as in Figs. 26, 30, and hence the difficulty of proper adjust- 
ment, maintenance and repairs. 

4. Stage Compressors, in which the air cylinders are 
compounded. The air end may be of the double-, triple-, or 
quadruple-stage type, according to the air pressure to be pro- 
duced. Stage compressors are now built by nearly all makers, 

, * American Machinist, Oct. i6, 1902, p. 1475. 



STRUCTURE AND OPERATION OF COMPRESSORS 



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CO^IPRESSED AIR PLANT 



and are the most important class.* (a) Straight-line form, as in 
(i). These have two-stage air ends, some with compound steam 
ends also. Fig. 6 shows a Xorwalk, and Figs. 7. 8 a SuUivan 
compressor, both ha\4ng compound steam and two- stage air cyl- 
inders. Fig. 5 is a two-stage compressor with simple steam end. 
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FiG. 2. — IngersoU-Rand Straight-Line Compressor, Class " FR-i ". 



air locomotives. Figs. 9. 10, and 11 show the latest t^-pe of 
IngersoU-Rand cross-compound, two-stage compressor, with pis- 
ton steam valves; similar compressors of smaller capacity are 
fitted with Meyer slide valves. Figs. 12, 13 are of two t\-pes of 
Laidlaw-Dunn-Gordon compressors, the former having Corliss 
inlet and poppet discharge valves for the air end, the latter 
being a newer design, with " feather " air valves. Fig. 14 shows 
the latest t>T)e of AUis-Chalmers compressor. For illustrations 
of 3- and 4-stage compressors, see Chap. XXVL 

* The Xorwalk Iron Works Co. was the pioneer in the field of stage compres- 
sion, having begun in 1 880-81 to build this t>-pe of compressor for ordinarj- semce. 



STRUCTURE AND OPERATION OF COMPRESSORS 



11 




12 



COMPRESSED AIR PLANT 





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STRUCTURE AND OPERATION OF COMPRESSORS- 



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STRUCTURE AND OPERATION OF COMPRESSORS 



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STRUCTURE AND OPERATION OF COMPRESSORS 



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COIMPRESSED AIR PLANT 




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20 co:mpressed air plant 

Among well-known British compressors of this t}7)e are 
those of Walker Bros., Wigan, and Robey & Co., Lincoln. The 
Walker compressor is horizontal, duplex, two-stage, with simple 
or compound steam end, and is made of capacities from i,ioo 
to 9,400 cu.ft. free air per min. Robey & Co. make high-speed, 
vertical (see f^ig. 30), as well as low-speed, horizontal com- 
pressors. The machines of both of these builders have large 
disk, or plate, air valves fChap. VII). The Riedler com- 
pressor, formerly made by Eraser (S: Chalmers, Erith, England, 
has been largely used for mining work. Eor its unique air 
valves, see Chap. IX. 

As based on structural characteristics, compressors may also 
be classified as: (a) Direct-driven by steam, electricity, or 
water-power — the motor end being directly connected with the 
air cylinders. Among water motors the bucket or impulse 
w^heels are best adapted to this service; (b) Belt-driven from 
independent motors: steam-engines, water-wheels, or electric 
motors. ' These are in com.mon use for mine and other service. 
Chain-driven and direct-connected compressors are also occasion- 
ally employed. (Eigs. 28, 29.) 

So-called " half-duplex " compressors consist of either the 
right- or left-hand half of a duplex compressor, an extended 
crank-shaft and out-board pillow-block being provided tem- 
porarily. If a comparatively small quantity of air is needed 
for a time — as during the development of a mine or the sinking 
of a shaft — a half -duplex may be installed at first, the second 
half being added later. The capacity is thus readily doubled at 
moderate cost. 

Comparison of Types of Compressor 

The straight-line compressor is largely employed for small 
or medium size plants, or for temporary ser\4ce. It is compact, 
strong, and self-contained, being carried on a single bed-frame 
and requiring a relatively inexpensive foundation. The floor 
space occupied is much less than for the duplex form. 

While useful for moderate air pressures and fairly constant 



STRUCTURE AND OPERATION OF COMPRESSORS 21 

loads, the straight-line compressor is not capable of operating 
with the steam economy essential in large plants: nor is it self- 
regulating at much less than, say, 40% of its full load. These 
compressors are usually made of capacities from the smallest 
up to 1,700 or 1,800 cu.ft. of free air per minute, the last-named 
sizes developing 275-300 H.P. 

The duplex compressor is preferable to the straight-line 
for large plants. It is better adapted to varying loads, arising 
from differences of air pressure, because the resistance is more 
uniformly distributed throughout the stroke. Having quarter- 
ing cranks it will run at very low speed without stopping on a 
center; it is self-regulating and capable of dealing economically 
with a range of load down to less than one-quarter or one- third 
of its normal. As a rule, the friction loss (total H.P. consumed 
by engine friction) is no greater and is often less than that of a 
straight-line of the same capacity. For large compressors, 
in good order, this loss may be not over 5-7%.* While these 
ngures are sometimes . equalled by the best straight-line com- 
pressors, the loss in the latter is generally higher. 

The Corliss type of engine is frequently used for driving 
large duplex compressors, as its valve gear is well adapted for 
dealing with variations of air pressure. Corliss valves have been 
largely employed in the past, for the air as well as the steam 
cyHnders. Since about 19 13 air valves of the thin- plate or 
" feather " type have been adopted by several well-known 
makers. (See Figs, i, 2, 3 and 24; also Chaps. VII, VIII and 
IX.) 

The foundation of the duplex compressor is necessarily more 
expensive than that of the straight-line, and to maintain per- 
fect alinement must be substantially built. Each pair of cyHn- 
ders are connected by trunk-frames or tie-bolts. A complete 
girder-frame (Figs, i, 2, 9, 12, 14) prevents any possibility of 
movement. The tandem steam and air cylinders on each side 

* See an article by J. Parke Charming, in Mines and Minerals, May, 1905, 
p. 475, containing the results of an efficiency test on a 300-H.P., compound, two- 
stage Nordberg Corliss compressor, at the Burra-Burra mine of the Tennessee 
Copper Co. Its total efficiency was found to be 78.1%. The horse-power con- 
sumed by friction was only 5.2% (see p. 149). 



22 COMPRESSED AIR PLANT 

are best placed far enough apart to prevent the same portion 
of the piston-rod from passing alternately into each stufhng-box, 
because: first, as the piston-rod is apt to wear differently in 
the two stufhng-boxes, it becomes difficult to keep them well 
packed and tight; second, the steam and air piston-rods are 
often in separate parts, coupled between the cyhnders. This is 
convenient in dismantling the compressor for repairs; also, 
the air valves, when of the poppet form and in the cylinder head, 
are more accessible. 

Compressors with Compound Steam Cylinders. The eco- 
nomic advantages of compounding the steam end are greater 
than in ordinary engines: first, because the conversion of 
power from one form to another is necessarily attended by loss, 
and should be conducted as efficiently as possible; second, 
because, as shown below, air compression involves unfavorable 
load conditions. A steam saving of, say, 20% may be readily 
attained by using a good condenser, thus getting the full 
expansive power out of the steam, and by avoiding loss of power 
due to imperfect speed regulation and consequent blowing off 
of air at the safety valve. 

Stage compressors. It is now recognized that even for 
pressures of 75-90 lbs., as commonly employed for machine 
drills, a saving in steam consumption can be realized by stage 
compression. In mountain regions, where so much mining 
is carried on, its advantages are still greater than at sea-level 
(Chap. XIII). The duplex form, with both steam and air ends 
compounded, exemplifies the highest t>'pe. There is no material 
increase in the number of moving parts, except valves; the 
greatest range of steam expansion is obtainable, because the 
work done in the air cylinders is more nearly equahzed, and the 
compressor may be made self-regulating over its entire range 
of load. Thermodynamically, the efficiency of stage com- 
pression depends largely on the proper use of water-jackets 
for the air cylinders, and the size and design of the intercoolers. 
(Chap. VI.) 

The Operation of Steam-driven Compressors involves con- 
ditions which do not obtain in ordinary steam engines (see Fig. 



STRUCTURE AND OPERATION OF COMPRESSORS 



23 



15). At the beginning of the stroke the air in front of the 
piston is approximately at atmospheric pressure. As the piston 
advances the pressure at first increases slowly, and then very 
rapidly to its maximum. The power developed in the steam 
cylinder, when working as usual with a cutoff, is in the reverse 
order. The initial steam pressure may be even lower than the 
final air pressure, though the mean effective pressure in the 
steam cylinder is greater than the mean effective in the air 
cylinder. For example, with an initial steam pressure of 60 lbs., 
air may be compressed to 80 lbs. or more. This is due to the use 
of heavy fly-wheels and reciprocating parts, which store up 
the surplus power in the early part of the stroke, and give it 
out toward the end. The consequent lack of smoothness in the 




Fig. 15. — Combined Air and Steam Cards, Straight-Line Compressor. 



running of compressors is especially noticeable in the simple 
straight-line type, which, when the air in the receiver is at 
maximum, often comes almost to a standstill and barely turns 
over the centers. It would thus appear that only a small ratio 
of expansion in the steam cyHnder could be employed; but the 
difficulty is met by the inertia of the fly-wheels, and the cylinders 
of even simple compressors can be proportioned for a steam 
cutoff at f to I stroke. In the duplex type, power and resistance 
are more nearly equalized, the most favorable distribution being 
attained in cross-compound, stage compressors. 

Steam Valves are of a number of forms: plain slide-valve, 
with or without Meyer adjustable cutoff, Corliss, drop valve, 
and balanced piston valve. 

In most simple straight-fine compressors the steam cylinder 



24 COMPRESSED AIR PLAXT 

has an adjustable cutoli valve (Fig. i). This valve (;2^S) is 
composed of two parts on a right- and left-hand threaded stem, 
and. by mo\'ing on top of the sHde valve, controls ports in the 
latter through which steam is admitted to the main ports. It 
is operated by a separate eccentric on the fly-wheel shaft, and by 
the hand-wheel (69), outside of the valve chest, may be regulated 
without stopping the compressor, according to the varpng 
receiver pressure. By manipulating this valve the engineer can 
prevent the compressor from stopping on a dead center, not- 
TN-ithstanding variations in air pressure (see also Fig. 12). 

Ihe Corliss valve is used by a number of makers, chiefly for 
large engines, both straight-hne and duplex, and is well adapted 
for compressor service (Figs. 6, 7, 8 and 14). The double-ported 
steam valves in Fig. 8 enable the compressor to run at a higher 
speed. 

Piston steam valves have recently been adopted by the 
Ingersoll-Rand Co. for its straight-hne and Imperial '* XPV " 
compressors (Figs. 2, 3, 9, 10). They are balanced, of the 
telescopic t\-pe. Their accompanying cutoft' valves, one for 
each end of the cylinder, are right- and left-hand threaded 
on the cutofl" valve stem, which telescopes through the main 
valve stem. Steam enters the interior of the main valve, 
and passes out through the ports near the ends, being exhausted 
by the ends of the valve. This design requires only two, instead 
of four, ports, thus reducing clearance and condensation surface. 
The main valve is in two parts; the ends, which separate live 
steam from the exhaust, are cast integral with each half, thus 
avoiding the need for steam-tight joints at these points. As 
steam exhausts past the ends, the valve covers and packing 
are under exhaust pressure only, thus decreasing leakage, 
and subjecting to low temperatures the exterior walls of the 
chest. This valve is well suited to high steam pressures and 
the use of superheating. 

Drop valves are adapted to large compressors working \\ath 
high steam pressure and superheat. This valve is a double- 
ported poppet, lifted by a cam and closed by a spring. It is 
nearly balanced, and, as dash-pots are not required, the running 



STRUCTURE AND OPERATION OF COMPRESSORS 25 

speed of the compressor may be higher than for Corhss engines. 
Both ports and steam passages are large. The point of admis- 
sion to the high-pressure valves is fixed, but an increase of air 
pressure beyond the predetermined point causes an earher 
cutoff; the reverse conditions, a later cutoff. The low-pressure 
admission valves and all exhaust valves are set to give constant 
points of admission and cutoff. 

Proportions of Cylinders. A short stroke is conducive to 
economy in compression, as well as the attainment of a proper 
rotative speed. It is of especial importance in simple straight- 
line compressors, because the power and resistance are then more 
nearly equalized. With a long stroke the piston would travel 
some distance under an increasing resistance; then, after the 
discharge valves open, it would complete its stroke under a 
uniform resistance, while adding nothing to the amount of 
useful work. But the loss of volumetric capacity due to piston 
clearance is less for a long than a short cylinder of the same 
diameter. In single-stage, slide-valve compressors, the usual 
ratio of stroke to diameter of steam cylinder is li : i or li : i. 
In some recent designs, the stroke and diameter are nearly 
equal, while in duplex Corliss compressors are found such varia- 
tions in the proportions of steam cyhnders as: 12X30 ins., 
14X42 ins., 20X42 ins., and 30X60 ins. 

The relative diameters of the air and steam cylinders de- 
pend obviously on the steam pressure carried and the air pressure 
to be produced. At mines there is usually but little variation in 
these conditions, except for operating compressed-air locomo- 
tives (Chap. XXVI). For rock-drills, the air pressure is gen- 
erally from 70-90 lbs. The appHcations of compressed air for 
manufacturing purposes have so multiplied that some builders 
furnish compressors to produce pressures of from 10-120 lbs. 
per sq. in. 

Compressors Driven by Gasolene Engines have recently 
been introduced. Fig. 16 shows a small portable outfit, with 
the compressor short-belted to the engine. It is useful for 
prospecting, quarrying, and general surface rock excavation. 
Figs. 17 and 18 show straight-line gasolene-driven compressors. 



26 COMPRESSED AIR PLANT 

They can be operated also on fuel oil and low-grade distillates. 
The power end of the compressor in Fig. 17 is of the valveless, 
two-cycle, low-compression type, without electric firing, and 
having the water supply and fuel injection to the combustion 
chamber positively governed. This compressor is built in 
four sizes, with air cylinder capacities (piston displacement) 
of 70-290 cu.ft. per min. (10-46 H.P.). It is also made in 
portable form, on wheels, and semi-portable, on skids. 

Compressors Driven by Water Power. Impulse wheels, 
as the Pelton, Knight, or Risdon, are best adapted to this 



Fig. 16. — TngersoU Rand IV-rtable Gasolene-Driven Compressor. 

service, the wheel being mounted directly on the crank-shaft, as 
in Figs. 19, 20. A plant somewhat similar in general layout 
was built by the Pelton Water Wheel Co., for the Alaska- 
Treadwell Gold Mining Co. Since the power developed is uni- 
form throughout the revolution of the wheel, the compressor 
should be duplex to equalize the resistance as far as possible, and 
the rim of the wheel is made extra heavy, to act as a fly-wheel. 
Fig. 21 shows a recent form of Pelton wheel, sectionalized for 
mule-back transport in Mexico, the hub being designed for 
shrinking on forged bands, after the wheel is in place on its 
shaft. The fly-wheel effect of small diameter water-wheels is 



STRUCTURE AND OPERATION OF COMPRESSORS 



27 




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COMPRESSED AIR PLANT 




STRUCTURE AND OPERATION OF COMPRESSORS 



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COMPRESSED AIR PLANT 







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STRUCTURE AND OPERATION OF COMPRESSORS 



31 



sometimes increased by heavy iron segments bolted on the 
spider. Water-driven compressors are also built by the Com- 
pressed Air Machinery Co., and others. 

For best efficiency, the peripheral velocity of an impulse 
wheel is theoretically one-half the velocity of the jet of water 
from the nozzle. Hence, high heads involve high peripheral 
velocities, and for a wheel of small diameter a belt-drive would 





Fig. 21. — Pelton Impulse Wheel. 



be required. But belting or gearing can generally be avoided, 
except for a turbine-wheel drive (loss of power in belt trans- 
mission is, say, 8-io%). An impulse wheel can generally be 
made of large enough diameter to run at a peripheral speed 
that will insure economical use of the water, while still having 
a sufficiently low rotative speed for direct-connected com- 
pressors. To accomplish this under very heavy heads, the 
>vheels are sometimes of great size, 



32 COMPRESSED AIR PLANT 

At the Morning Mine, near Mullan, Idaho, is a large water- 
driven two-stage compressor. There are 4 cylinders, a high- 
and a low-pressure being set tandem on each side of a set of 3 
Pelton wheels, mounted on the crank-shaft. A large volume of 
water, under a head of 140 ft., is delivered through 8,000 ft. of 
flume and 400 ft. of pressure pipe^ driving two 12-ft. wheels. Two 
other streams, piped respectively ij and i mile, produce heads 
of 1,140 and 1,420 ft.; these drive a 33-ft. Pelton wheel, placed on 
the crank-shaft between the smaller wheels. The central wheel 
is driven by separate jets from the high-pressure lines, and on 
account of their difference in head, the diameter of this wheel 
is a mean between the diameters theoretically necessary for 
obtaining a peripheral velocity properly proportioned to each 
head. The actual mean peripheral speed is 8,000 ft. per minute. 
To control the water under these great heads (pressures, about 
490 and 610 lbs. per sq. in.), slow-acting gate valves are provided, 
with by-passes for starting and stopping. For stopping the 
compressor quickly the nozzles can be deflected clear of the 
wheel. 

Each pair of cylinders are ^^^ and 18 ins.X42-in. stroke; 
piston speed, 560 ft. The low-pressure cylinders compress to 
about 30 lbs., the high-pressure to 90 lbs. Inter- and after- 
coolers are placed in the tail-race of the smaller wheels. A 
positive valve-motion is employed for both inlet and discharge 
valves, which are of the Corliss type. On each side, parallel 
to the center line of the compressor and geared to the crank- 
shaft, is a long shaft, geared to which are short shafts carrying 
the valve eccentrics. As the discharge valves must open when 
the pistons are moving at nearly maximum velocity (800 ft. 
per min.), an auxiliary dash-pot allows them to open freely 
under the cylinder pressure, the positive eccentric motion closing 
them. 

Indicator cards show this compressor to be highly efficient. 
An average of a number of cards gives mean pressures of: low- 
pressure cylinder, 17.86 lbs.; high-pressure, 41.14 lbs.; com- 
bined, 30.46 lbs. The mean theoretical adiabatic and isothermal 
pressures, corresponding to the combined mean are, respectively, 



STRUCTURE AND OPERATION OF COMPRESSORS 33 

36.94 and 28.5 lbs. During the tests the observed temperatures 
were: coohng water, 38°; air at discharge from low-pressure 
cylinder, 135°; at high-pressure inlet, 46°; high-pressure dis- 
charge, 140°; on leaving the after-cooler, 62°. Mean atmos- 
pheric temperature, 55° and of the cooling water 38°. 

If there is a sufficient volume of water, impulse wheels may 
be used with low heads, by introducing multiple nozzles, directed 
tangentially at two or more points of the periphery of the wheel. 
To prevent water from splashing over the compressor, the wheel 
is enclosed in a tight wooden or iron casing. The flow of water 
may be regulated by an ordinary gate valve; but if the head be 
great a special slow-moving gate must be used (as noted above), 
to avoid danger of rupturing the pressure pipe in case the com- 
pressor is suddenly stopped. Turbines are obviously not so 
well adapted for operating compressors as impulse wheels. 

Nozzles are now usually of the " needle " type (Fig. 22). 
The position of the needle is adjusted by a hand- wheel and 
gearing for delivering the requisite volume of water, the needle 
being shaped to form a solid stream against the buckets of the 
wheel. The cut shows also an automatic cutoff governor, 
controlled by air pressure from the compressor receiver. An 
excess of receiver pressure, admitted to the small cylinder a, 
causes arm B to move backward, thus raising hood c in front of 
the nozzle and deflecting the water to the tail race, instead 
of allowing it to impinge on the wheel. When receiver pressure 
returns to normal, the hood automatically moves downward, 
out of hne of the jet. This regulates the speed of the wheel 
(and also of the compressor), but does not economize water 
consumption by checking the nozzle flow. To save water, there 
are several devices for automatically regulating the opening 
of a valve in the pipe from penstock to nozzle, and so adjusting 
the stream impinging on the wheel. 

In another design, the nozzle itself is deflected out of line 
with the buckets on the wheel. 

Belt-driven Compressors. The fly-wheel is replaced by a 
belt-wheel with a heavy rim to give it sufficient weight. Figs. 
23, 24 show straight-hne and duplex, two-stage compressors. 



34 



COMPRESSED AIR PLANT 



Fig. 25 is a compact form of small two-stage compressor, 
suitable for underground installation, or where floor-space is 
limited. The Sullivan Machinery Co. also builds small belt- 
driven, single-stage compressors for surface and underground 
use, and a portable- gasolene-driven compressor. For under- 
ground service, the '' W K " and " W K-2 " types are mounted 




Fig. 22. — Needle Nozzle for Pelton Wheel. 



on a light, self-propelling truck; they are made in 4 sizes, 
capacity 98-258 cu.ft. of free air per min. 

Fig. 26 shows a well-known English compressor made in a 
number of sizes for both steam and belt-drive. The two-stage 
machine has a differential piston, working in a single cylinder. 
Air is admitted to both the low- and high-pressure parts of the 
cylinder by one large piston valve, which controls absolutely 
the periods of admission and also the end of the discharge 



STRUCTURE AND OPERATION OF COMPRESSORS 



35 



p^^ 




36 



COMPRESSED AIR PLANT 




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STRUCTURE AND OPERATION OF COMPRESSORS 



37 



periods. The actual discharge is through groups of poppet 
valves. 

Fig. 27 is the high-pressure side of a recent duplex, two- 
stage compressor. Power may be derived from an engine 
already installed for other purposes, or from a water-wheel, 



Water Outlet-E 
Air Outlet ^^ 




Air Inlet 
Water Inlet 



Fig. 25. — Sullivan Belt-Driven, Two-Stage, Angle Compressor, Class " WJ-3 " 

(1917)- 



electric motor, or gasolene engine. Small portable motor- 
driven compressors mounted on wheels (like Fig. 16) are also 
to be had. 

Some builders employ a " silent-chain " drive, when it is 
desired to place the motor close to the compressor and on the 



38 



COMPRESSED AIR PLANT 



same bed-frame, and at the same time to avoid the use of gear- 
ing (Fig. 28). It has high efhciency (about 95%), and will 
transmit up to, say, 200 H.P. 

Compressor Direct-connected to Electric Motor. Although 
a belt-drive is preferable to gearing, at least for a compressor 
erected on the surface, geared electric-driven sets are sometimes 




Fig. 26. — Alley and MacClellan Vertical, Two-Stage, Belt-Driven Compressor 
(from paper by G. Blake Walker, Trans. y Midland Inst., Min., Civil and Mech. 
Engs., Jan. 21, 1913). 



used, a spur-gear on the crank-shaft engaging with a pinion on 
the armature. This design has been adopted for large plants, 
as, for example, at a two-stage installation of the Compania de 
Penoles, Mexico. By giving sufficient diameter and weight to 
the spur-wheel, it not only produces a low piston speed but 
serves also as a fly-wheel. Rawhide pinions are desirable to 



STRUCTURE AND OPERATION OF COMPRESSORS 



39 




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COMPRESSED AIR PLANT 



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reduce noise. The small, high-speed Chris tensen compressor 
is well adapted for gearing to a motor, thus forming a compact 
machine where Ughtness or portabihty is essential. Fig. 29 
shows a direct-connected, duplex compressor. The frame is 
inclosed and all parts are self -oiling, except the piston and 
cyhnder. The crank-pit contains the oil supply. The oil is 
picked up by the edge of the crank disk, and taken off at the top 




Fig. 28. — IngersoU-Rand Compressor, Chain-Driven from an Electric Motor. 
Class "NE-i," i2i"Xi2" Cylinder. 

by a scraper; part goes to a distributing tank, from which small 
pipes lead to the crosshead and guides. The main bearing 
and crank-pin are oiled direct from the scraper by a projecting 
trough on each side. For the bearing the oil is led to a channel 
in the cap, and thence through a series of holes drilled in the 
bearing liner. Some of this oil goes to the collar of the eccentric 
hub, from which it is carried to the face of the eccentric through 
two holes. As the eccentric is closed, the surplus oil returns 
to the crank-pit. Thus the oil-feed is proportioned to the speed 



Intercooler 




Air Cylinder 
Regulator 

or Unloader, 
inserted in 

Intake Pipe 



Fig. 29. — Ingersoll-Rand, Direct-Connected, E 
Type 10. Longitudinal Secti< 




Driven, " Imperial " Duplex Compressor, 
Low-Pressure Cylinder. 



To face page 40. 



?:.j; 



STRUCTURE AND OPERATION OF COMPRESSORS 



41 



of the compressor, ceasing entirely when the compressor is 
stopped, and results in a material saving. * 

Fig. 30 is a well-known English direct-connected compressor. 
It has plate air valves (Chap. VII), forced lubrication, and large 
water-jackets. 

These compressors are driven by direct-current, induction 
or synchronous motors, the rotors of which are of large diametefr, 




Fig. 30. — Robey & Co. High-Speed, Vertical, Two^Stage, Motor-Driven 

Compressor. 

as shown, to produce a proper relation between peripheral and 
rotative speeds. Induction motors are good, as they run econom- 
ically under wide variations of load. 

Under proper conditions an electric-driven compressor may 
be erected underground, near the point of application of the air 
power. A difficulty in operating compressors underground 

* Similar self-oiling (" flood and splash ") systems are used for compressors 
shown in Figs, i, 4, 9, 17, 18, 23, and 24. 



42 COMPRESSED AIR PLANT 

is in obtaining cooling water of good quality and in sufficient 
quantity. Mine water usually contains enough sediment to 
foul the inner surfaces of the jackets and intercooler. Also, 
the heated water must be cooled before reuse; for example, by 
pumping it through a worm pipe laid in the sump. Hence, the 
volume of water entering and pumped from the sump is impor- 
tant. While a small compressor (of say 300 cu.ft. free air per 
min.) might be successful underground, a large one would be 
out of the question. 

Turbo-compressors. Based on the principle of his steam 
turbine, Parsons designed a type of low-pressure compressor, 
useful as a blowing engine. From this, Rateau developed the 
turbo-compressor, now built by the Westinghouse Electric Co., 
Ingersoll-Rand Co., and other makers in Europe and the United 
States. They satisfactorily produce pressures suitable for mine 
service. For pressures above 2 or 3 atmospheres, the com- 
pression is done in stages, the number of impellers per stage 
depending on the required pressure. Fig. 31 shows a 2-stage 
turbo, with 10 impellers per stage. 

The rapid rotation of the impellers (rarely less than 3,000 
R.P.M.) imparts to the air a velocity of 300 ft. or more per 
second. The velocity at which the air issues from each impeller 
is converted into head, or pressure, as the air passes into the 
larger passages leading to the next impeller, each adding an 
increment of pressure. This increment is normally 2-2 1 lbs., 
sometimes 3-4 lbs., or even more. 

Water jackets are applied to the spaces between the impeller 
casings, diffusers and passages. In the first impellers of the 
series, where the pressure is low, the temperature rises rapidly, 
notwithstanding the jacketing; but, as the density of the air 
increases, cooling becomes more effective, and the total isother- 
mal efficiency is good. The volume of cooling water required, 
at 70° F., is roughly 165 gals, per min. per 1,000 H.P. Some 
engineers estimate 3,000 gals, per hr. per 1,000 cu.ft. free air 
compressed per min. 

Examples. On the Rand, South Africa, a number of large 
turbos are operated by the Victoria Falls and Transvaal Power 



STRUCTURE AND OPERATION OF COMPRESSORS 



43 



Co. Power is transmitted from 
central stations through more 
than 1 8 miles of piping to 17 
mines. The last compressor in- 
stalled has a capacity of 58,800 
cu.ft. free air per min. to 140-170 
lbs. gage, requiring 12,000-13,000 
H.P. It is driven by a steam 
turbine using superheated steam 
at 170 lbs.; speed, 3,000 R.P.M. 
There are 3 sets of impellers (3 
stages). The same company has 
12 4,000-H.P. turbos in opera- 
tion, each delivering 20,000- 
23,000 cu.ft. free air per min., 
compressed to 135-170 lbs., and 
driven by synchronous motors. 
They are 4-stage : 2 low-pressure, 
I intermediate and i high- 
pressure cylinders; speed, 3,000 
R.P.M. Efficiency (referred to 
isothermal compression), 67.5%. 

Fig. 31 shows a 2-stage turbo, 
driven by a steam turbine; 
capacity, 7,500 cu.ft. per min. to 
80 lbs.; brake H.P., 1,480, at 
4,000 R.P.M. Each stage com- 
prises 10 impellers, similar to 
those in Fig. 32 (for details, see 
Trans. Instn. Min. Engs., Eng- 
land, Vol. 45). 

Another turbo, driven at 4,200 
R.P.M. by a i,ooo-H.P. mixed- 
pressure turbine, compresses 4,400 
cu.ft. air per min. to 99 lbs. 
Consumption of exhaust steam 
(taken at 15.6 lbs. and discharged 



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6 



^ 

^ 



o 



44 COMPRESSED AIR PLANT 

at 1. 14 lb.) was 7.98 lbs. per 100 cu.ft. free air; with this were 
used 5.05 lbs. hve steam at 85 lbs. 

The final temperature for a delivery pressure of 85-115 
lbs. is say 130°-! 70° F., depending on temperature of the cooling 
water. In one case, with cooHng water at 50° F., and a delivery 
pressure of 120 lbs., the final temperature was 136° F. 

Turbos compressing 5,000-12,000 cu.ft. free air per min. 
to 85-120 lbs. should run about 3,800 R.P.M.; smaller machines, 
down to 2,500 cu.ft. at 70-85 lbs., 1,750 cu.ft. at 60 lbs., and 
600 cu.ft. at 25 lbs., should be designed for 4,200-5,000 R.P.M. 
(G. Blake Walker, Trans. Instn. Min. Engs., England, Vol. 44, 
pp. 629-689). 



Fig. 32. — Series of Impellers, Westinghouse Turbo-Compressor. 

Fig. 33 shows results of tests on the turbo at the New Huck- 
nall colliery, England. Turbos exceeding 6,000 cu.ft. free air 
capacity should give 70-80% efficiency (referred to isothermal) ; 
between 3,000 and 6,000 cu.ft., 65-70% efficiency at full load, 
and under 50% at less than half load. 

Field of Use. An important use for turbo-compressors 
is in furnishing blast for smelting furnaces, Bessemer converters, 
and other metallurgical service; air pressure being, say, 5-35 
or 40 lbs. The chief difference between blowers and high- 
pressure turbos is in the number of impellers and provision for 
deahng with the heat of compression. The efficient governors 
of some recent turbos give a nearly constant pressure, even with 
quite -svide fluctuations in the volume of air used. For example, 
a 5,000-cu.ft. Thomson-Houston turbo, running normally at 
4,600 R.P.M. , showed a variation of only 2 or 3 lbs. for a range 
in output from 1,500 to 6,000 cu.ft. free air per min. 



STRUCTURE AND OPERATION OF COMPRESSORS 



45 



In point of first cost, it is generally not advisable to use turbos 
for capacities smaller than 2,200-2,500 cu.ft. free air per min., 
nor for pressures exceeding 70-85 lbs.; for smaller output or 
higher pressure, reciprocating compressors are best. 

Note. — It is not practicable, in a book that is not a trade publication, to describe all 
the makes of air compressors. Those which have been instanced illustrate the chief features 
of design. The same is true regarding the descriptions of air valves, etc., in Chaps. VII, 
VIII, and IX. It must not be understood '^hat the compressors referred to are consider,ed 
the only good ones, nor that the author, b^" omitting to mention and insert cuts of all com- 
pressors, desires to discriminate against those that are less well known. Most of the Euro- 
pean compressors, including many excellent machines, are omitted altogether, though refer- 
ences to the valve motions of some of them will be found under the appropriate heads. 



35,000 



80,000 S 



25,000 



o) 20,000 o 

t3 





S p. 
i 15,000 -a 
CO 3 

g i 



10,000 



5,000 1 20 



10 



Guarantee Figures— 






_o 












.*-^ 


r 




Guarantee and Test Figures (@) 










■ 


f^ 


y 
















^ 










A 
























^^>^' 




*e^ ^ 


X 




























y 


><^ 


^ 


;^vb 








/ 


^^- 














/ 


/^ 






^ 




5^ 


^ 


^ 


r^ 












< 






,j^^^' 




b 






^ 
















>^ 


^ 


<"^ 

We^ 


Jteafv, 
























^ 


r"^ 








l^erj^ 


'""Co 


^io Fi 


et of 


lir 






Lives 


team 


30 Po 


mdsi 


Guarantee bondi 
er Square Inch ( 


;ions; 
Tauge 
iches- 
bsolu 
ches 


27J^ 


■nches 


1 of Y£ 


icuum, 






Exh 


aust S 


team 


16 Po 


mdsi 


er Sq 
Bare 


aarel 
meter 


ach A 
, 30Ir 


;e,2T}' 


r Inch 


esof 


^''acuu 


Ul, 





































































































1,000 2,000 3,000 4,000 5,000 6,000 7,000 

Output, Cubic Feet per Minute of Free Air, at 30 Inches Barometer and CO°Falir 



8,000 



Fig. 33. — Results of Test of British-Westinghouse Rateau Mixed-Pressure Tur- 
bine and Air Compressor. Speed, 4,000 R.P.M.; output, 7,500 cu.ft. free 
air per min., at 80 lbs. gage. Test made by W. F. Mylan (Trans. Instn. 
Min. Engs., England, Vol. 45, pp. 245-264). 



Following is a detailed classification of the compressor 
practice of the Ingersoll-Rand Co. Attention is called to the 
note regarding the two-stage, straight-line compressors of this 
make. Several other builders (for example, the Sullivan 
Machinery Co. and the Nor walk Iron Works Co.) furnish two- 



46 



COMPRESSED AIR PLANT 



stage, straight-line compressors 
used for rock-drills, pumps, etc. 



for the ordinary pressures 



STEAM DRIVEN 



Straight-line. 



Duplex, small and 
medium capacity. 

Duplex, Large 
large capacity. 

Turbo-compressors. 



Single-stage, double 

acting. 



(Direct-connected to steam engine. 
For long belting to steam engine. 
For short belting to steam engine. 
Note. Two-stage, straight-line compressors for steam operation are 

built for high-pressure work only. 
Single-stage, double 
acting, portable. - 



(Single-stage. 
Two-stage. 

j Single-stage. 
\ Two-stage. 



Direct-connected to steam engine. 



Direct-connected to steam engine. 
For long belting to steam engine. 
For short belting to steam engine. 

Direct-connected to steam engine. 

Direct-connected to steam turbine. 



POWER DRIVEN 



Vertical. 



Straight-line. 



Duplex, small and 
medium capacity. 

Duplex, large 
capacity. 



Single-stage, single-acting, 
stationary. 

Duplex single-stage, 
single-acting, portable. 



For long belting to line shaft, gaso- 
lene engine, or electric motor. 
For short belt drive. 

For short belt drive. 



Single-stage, double- 
acting. 



For long belt drive. 
For short belt drive. 
[ For chain drive. 
Note. — Two-stage compressors of this type are built for high- 
pressure work only. 

For short belt drive. 
For chain drive. 



Single-stage, double- 
acting, portable. 



Single-stage. 
Two-stage. 

Single-stage. 
Two-stage. 



J For long belt drive. 
\ For short belt drive. 

For long belt drive. 
Direct-connected to electric motor, 

gasolene, gas, or oil engine, or 

water wheel. 



The following alphabetical list, while incomplete, comprises 
the names of most of the American compressor-builders. 



AlHs-Chalmers Manufacturing Co. 
American Air Compressor Works. 
Bury Compressor Co. 
Chicago Pneumatic Tool Co. 
Clayton Air Compressor Works. 
Compressed Air Machinery Co. 
Ingersoll-Rand Co. 
Laidlaw-Dunn-Gordon Co. 



New York Air Compressor Co. 

Nordberg Manufacturing Co. 

Norwalk Iron Works Co. 

Rix Compressor and Drill Co. 

Sullivan Machinery Co. 

Vulcan Iron Works. 

Worthington Pump and Machine Co. 



CHAPTER III 
THE COMPRESSION OF AIR 

In the production and use of compressed air occur serious 
losses, which to a large extent are unavoidable. Even in the 
best compressors the efficiency, or ratio of the force stored up 
in the compressed air to the work expended in compressing it, 
rarely exceeds 75% and often falls below 60%. To understand 
the causes of these losses it is necessary to study the principles 
of air compression. This study is advisable, also, before pro- 
ceeding to a description of the air end of the compressor. Several 
definitions follow: 

** Free air " is air at normal atmospheric pressure, as taken 
into the compressor cylinder. But since atmospheric pressure 
varies with the altitude above sea-level, and with the barometric 
reading at any particular time or place, the expression ^' free 
air " has no precise signification, with respect to the pressure, 
volume, and temperature of the air. At sea-level it is in reality 
" compressed air," at the normal atmospheric pressure of 14.7 
lbs. per sq. in. As commonly employed the term means air at 
sea-level pressure, and at a temperature of 60° F. 

The absolute pressure of air is measured from zero, and is 
equal to the assumed (or observed) atmospheric pressure plus 
gage pressure; ordinary gages register pressures in lbs. per 
sq. in. above atmospheric pressure. 

Absolute temperature is the temperature as measured from 
the " absolute zero " point, which is 491.4° F. below the freezing- 
point of water, or say 459° below zero F. Thus 60° F. of thermo- 
metric temperature is equivalent to an absolute temperature of 
459°+6o° = 5i9° F. 

Two fundamental laws govern the behavior of a perfect gas 
when undergoing compression, which for practical purposes 

47 



48 



COMPRESSED AIR PLAXT 



apply also to atmospheric air. In discussing the problems of 
air compression, all the relations between volume, pressure, 
and temperature may be expressed in accordance with these 
laws. The first law (Boyle's) is: At constant temiperature 
the volume occupied by a given weight of air varies inversely 
as the pressure. This is expressed by: 



P' V 
P\ = P'\ ' = constant ; or — - = =^7 ; 

P V 



in which 



V = volume of the given weight of air (or gas) at the freezing- 
point and at a pressure P (V usually being taken as the volume 
in cu.ft. occupied by i lb. of air); V'= volume of the same 
weight of air at the same temperature and at any pressure P' 
(the pressures being absolute pressures). 

For example, to compress a quantity of atmospheric air at 
constant temperature to 0.147 of its original volume (atmos- 
pheric pressure being 14.7 lbs.), requires a pressure of 100 lbs. 
per sq. in. ; when compressed to 0.074 of its original volume, the 
pressure required is 200 lbs., and so on. 

Table I — (D. K. Clark) 



Tempera- 
ture. 
Deg. F. 


Weight of 

I Cu.ft. in 

Lbs. 


Volume of ' 
I Lb. in 
Cu.ft. 


Tempera- 
ture. 
Deg. F. 


Weight of 

I Cu.ft. in 

Lbs. 


Volume of 
I Lb. in 

Cu.ft. 





.0863 


11.582 


no 


.0607 


14 345 


10 


.0845 


11.834 


120 


.0685 


14 


596 


20 


.0827 


12.085 


130 


.0674 


14 


847 


30 


.0811 


12.336 


140 


.0662 


15 


098 


32 


.0807' 


12.386 1 


! 150 


.0651 


15 


350 


40 


.0794 


12.587 


160 


.0641 


15 


601 


50 


.0779 


12.838 , 


170 


.0631 


15 


852 


60 


.0764 


13.089 


180 


.0621 


16 


103 


62 


.0761 


13 141 


190 


.0612 


16 


354 


70 


•0750 


13-340 


200 


.0602 


16 


605 


80 


.0736 


13 592 


210 


-0593 


16 


856 


90 


.0722 


13-843 


212 


-0591 


16 


907 


100 


.0710 


14.094 ' 









Table I shows the weight and volume of dry air, at tem- 
peratures from o°-2i2° F., and at atmospheric pressure. 



THE COMPRESSION OF AIR 49 

The production and use of compressed air are not governed 
solely by Boyle's law. During compression heat is generated, 
and when the air is allowed to re-expand to its original volume 
this heat is taken up. If there is no transference of heat, the 
internal work, manifested by the increase of temperature, is 
independent of the time occupied by compression. This con- 
ditionals expressed by the second law, that of Charles and Gay- 
Lussac: When under constant pressure, the volume of a gas 
expands or contracts for each degree rise or fall of temperature, 
from freezing to boiling, by a constant fraction of the volume 
which is occupied at the freezing-point. Stated in another way, 
the volume of a gas under constant pressure is nearly pro- 
portional to the absolute temperature. The equation may be 
written: V' = V(iH-a/°). The complete relations between pres- 
sure, volume, and temperature are expressed by the equation: 
P'V' = P'V(i+aO, in which P' and V represent the pressure 
and volume of a given weight of air (or gas) at f F. above the 
freezing-point, V is the volume of the same quantity at the 
freezing-point, and a the coefficient of expansion of air, which is 
practically constant and is very nearly 4-g^Y on the Fahrenheit 
scale. Hence, for a rise in temperature of 1° F., the volume of 
air increases by 49T of the volume occupied at the freezing-point, 
under the same pressure (491° F. being the absolute temperature 
below freezing). 

The practical application of this law is that the heat generated 
reacts upon the air under compression, and increases the pressure 
due merely to the reduction in volume. By cooling the com- 
pressed air to its original temperature the pressure would be 
reduced to the normal amount, according to the first law. That 
is, the heat produced by compressing a given volume of air 
corresponds in degree to the cold resulting from the re-expansion 
of the same quantity of air to its original volume and 
pressure. 

Two other statements may be deduced from what precedes: 
I. Under constant pressure the volume of air varies directly as 
the absolute temperature; 2. The volume being constant, t)i§ 
absolute pressiire varies directly as the absolute temperature. 



50 COMPRESSED AIR PLANT 

The first of these statements is expressed thus: - 

V V 

— = -7- = constant, 

t t' ' . 

in which / and t' are absolute temperatures; whence, from 
Boyle's law: 

P — = P -7 = constant. 

For convenience, this constant is commonly denoted by R, and 

. . . PV 

the general equation is written PV = R/, or — = R. 

1/ 

The value of R is found as follows : Since for a given weight of 
gas or air the density, D, is inversely proportionate to the volume, 

V = , .08073 being the weight in lbs. of i cu.ft. of dry air, 

.08073 

at sea-level pressure (14.7 lbs.) and 32° F. The normal atmos- 
pheric pressure per sq. ft. = 14.7X144 = 2,116.8 lbs If, there- 
fore, I cu.ft. be expanded by the application of heat to a vol- 
ume of 2 cu.ft., the work done against atmospheric pressure, 

per lb. of air, will be ^ = 26,220 ft. lbs. To double the 

. 08073 

volume, according to Charles' law, would require the expenditure 
of 491.4° F. of heat; whence, in raising the temperature 1° F., 
the external work done by expansion is : 

PV 26220 ^ 

t 491.4 

The heat generated during compression and corresponding to 
different pressures is shown in Table II, the volume at normal 
atmospheric pressure being i, at a temperature of 60° F. This 
table shows that the rate of increase of temperature is not 
uniform, but diminishes as the pressure rises. Thus, from 
I to 2 atmospheres the increase is 115.8°; from 2 to 3, 79.3°; 
from 3 to 4 atmospheres, 62.3°, etc. The quantity of heat 



THE COMPRESSION OF AIR 



51 



generated during compression may be calculated by the fol- 
lowing formula: 

Q = ^— XNap. log.--, m which 

Q = quantity of heat in thermal units; 

R = constant = 96.03 7 (French unit) or 53.37 (English unit); 

/ = absolute final temperature in degrees, corresponding to 

V (centigrade scale for French and Fahrenheit for English units) ; 

J = value of one thermal unit = 1,400 ft. lbs. (or 778 ft. lbs. 
if English units be used); 

V and V^ = volumes of air, cu.ft., at beginning and end of 
compression. 



Table II 



Pressure in 
Atmospheres. 


Absolute Pressures, 

Lbs. per Sq.in. 

above Vacuum. 


Volumes in 
Cu.ft.. Adia- 
batic Compres- 
sion. 


Final Tempera- 
tures, Deg. F. 


Corresponding 

Increases of 
Temperature. 


I.OC 


14.70 


I. 000 


60.0 


00.0 


I-2S 


18.37 


0.854 


94 


.8 


34 


.8 


150 


22.05 


0.750 


124 


9 


64 


•9 


2.00 


29.40 


0.612 


175 


8 


115 


8 


2.50 


36.70 


0.522 


218 


3 


158 


3 


3.00 


44.10 


0.459 


255 


I 


195 


I 


350 


51 40 


O.411 


287 


8 


227 


8 


4.00 


58.80 


0.374 


317 


4 


257 


4 


5.00 


73-50 


0.319 


369 


4 


309 


4 


6.00 


88.20 


0.281 


414 


5 


354 


5 


7.00 


102.90 


0.252 


454 


5 


. 394 


5 


8.00 


1 1 7 . 60 


0. 229 


490. 


6 


430 


6 


9.00 


132.30 


0.2II 


523. 


7 


463. 


4 


IO.CX3 


147.00 


0.195 


554- 





494- 





15.00 

1 


220.50 


0.147 


681. 





621 . 






As the rise in temperature due to compression is proportional 
to the ratio of the final absolute pressure to the initial absolute 
pressure, the quantity of heat generated during compression to 
any given pressure, and the consequent work done, is greater at 
high altitudes than at sea-level. 



52 



COMPRESSED AIR PLANT 



The above conclusions are illustrated by the diagram, Fig. 
34.* It is, in reality, two diagrams, combined to save space. 
Firsts beginning at the lower left-hand corner, and curving up- 



"Volumes 



■9 12 



611 
•Sio 

m 
a> 

IH 

3 Q 
CO a 

CO 

(U 

u 
P^ 8 

•7 

•6 

5 

14 

S 

























\ 




\ 


















\ 




\ 


















\ 






\ 
















\ 






\ 


















\ 


\ 


















\ 


V 




















w 




















^. 


V 
















\9 








































V 








c5 


















\ 


i/ 














\ 




\ 


V 














^ 




\ 


/ 



















\ , 


\ 


















nA 


y 
















/ 


'\ 


\^ 












^' 


^ 


/" 


^ 


\^ 


\\ 






,^ 






^^ 








^ 


\ 



294.0 

279.3 

261.5 

249.9 

236.2 

220.5 

205.8 

191.1 

176.4 ^ 

161.7 I 

147.0. g 
hi 
■3 
m 

132.3 g 

117.6 



ClO 

a 
102.9 ^ 

73.6 
58.8 
44.1 
29.1 
1L7 
0.0 



C* iH 



Temperature, Fah. 

Fig. 34. 

ward, are the adiabatic and isothermal compression lines. 
Their intersections with the horizontal and vertical lines give 
the volumes of the unit of air when subjected to any given 

* Taken from " Compressed Air Production," by W. L. Saunders, several slight 
corrections having been made in the adiabatic and isothermal lines. 



THE COMPRESSION OF AIR 53 

pressure, by reading the figures at the top, and right- or left- 
hand margin of the diagram. The initial volume is tkken as i, 
and the spaces between the vertical lines are each one-tenth. 
The resulting volume is independent of the initial temperature 
of the air. The corresponding pressure may be read in terms 
of either gage or atmospheric pressure. Second, beginning at 
the lower right-hand corner of the diagram, and rising toward 
the left, are the lines of temperature, the assumed initial tem- 
peratures being o°, 60° and 100° F. The temperature cor- 
responding to any given pressure is read on the lower margin. 
It should be observed that these temperature curves are those of 
adiabatic compression. 

It follows from the above that if the temperatuie of the air 
rises during compression an increase of work ensues. 

Isothermal and Adiabatic Compression. In accordance with 
the laws already stated, air may be compressed in two ways: 

Isothermal Compression. — The temperature is kept constant 
during compression, the heat generated being abstracted as fast 
as it is produced. In this case the pressure of the air varies 
according to the equation PV = P'V, and the compression 
curve of an indicator diagram is an isothermal curve. 

Adiabatic Compression. — The temperature may be allowed 
to rise unchecked during the period of compression, as it will 
when there is no transference of heat by radiation or coohng 
devices. The rise in temperature increases the pressure due 
to reduction of volume only. Thus, the pressure rises faster 

... P' V 

than the volume dimmishes, and — becomes greater than —,. 

This relation is determined by considering the specific heats 
of air at constant pressure and at constant volume. The 
specific heat of any gas or vapor at constant pressure, Cp, is the 
quantity of heat required to raise the temperature of i lb. of 
the gas 1° F., the pressure being unchanged. The specific 
heat at constant volume, C^, is the quantity of heat required 
to raise the temperature of the gas 1° F., the volume being 
unchanged. Regnault found that for air 0^ = 0.2375 and 
Cp = o.i689. Cp is the greater, because external work is done 



54 COMPRESSED AIR PLANT 

during a change of temperature, if the pressure be constant 
and the air free to expand; under constant volume, no work 
is done upon external resistances. When, as in adiabatic com- 
pression the heat generated reacts on the air under compres- 

sion and increases the value of ^5-, to maintain the equation 

V . . 

:rp must be increased by an amount equivalent to the external 

work performed. The specific heats may be expressed in heat 
units, as above; or, by multiplying them by the mechanical 
equivalent of a heat unit (778 ft. -lbs. = J), they are given in 
terms of ft. -lbs. and are then denoted by K; that is, 

]Cp = Kp and JCi = Kp. 

Since Kp = CpX778 = 184.77, a,nd Kp = CpX 778 = 131.4: 

Kp 184.77 



Kr 131. 4 



= 1 . 406. 



This ratio is commonly denoted by n, and is the exponent of the 

power to which ;rp must be raised to make it equal to 15-*; 

n may also be expressed as equal to the ratio of the specific 
heats at constant pressure and volume: 

Cp 0-2375 

77- = 7^ = I . 406 = ft. 

Lv 0.1689 
The general equation for adiabatic compression is therefore: 

PV'^ = P'V''^ or p-=(y7 

Work of Compressors without Clearance. Isothermal Com- 
pression. The work done by a compressor without clearance, 
and using isothermal compression, is represented by the area 

* A statement of the proof of this deduction is unnecessary here; it is given in 
several books on Thermodynamics, for example, in Perry's work on the Steam 
Engine, p. 333. 



THE COMPRESSION OF AIR 



5(1 



under the compression curve (Fig. 35 j. Let AB be an isother- 
mal curve, AD representing any volume V of free air, and BC 
the volume V^, to which this quantity of air is compressed; 
the corresponding absolute pressures being respectively P and P'. 
The curve is an equilateral hyperbola, and the work W (repre- 
sented by the area ABCD)=Wi+W2 — W3, in which 

Wi = area under AB = fp^V. 

W2 = area under BC = P'V, representing the work of 
expelling the air from the cylinder. 

W3 = area under DA = PV, representing the negative 
work done by atmospheric pressure on the 
suction or intake side of the piston. 



I 
I 
I 
I 
I 
I 
I 
I 
« 

I 
I 
I 
I 
I 

I — 
I ^ 
I p 



p 
















c 




B 


E 


Note; V=BCorEC, 
according as the com- 
pression is isothermal 
or adiabatic. 






<r 


— Vt- 


-A 


A 










\ 


V 






D 








^^^^^^Si^^ 




\ 




^__ 














V 






— V 



Fig. 35. — Reference Diagram. 



Since PV = P'V', W2 and W3 cancel, so that the algebraic sum of 
W = Wi+W2-W3 = JpJV . .... (i) 



/ir/ 



To integrate this expression, substitute for P its equivalent 



PT 

V • 



W= PV^ = PV 



V 



Integrating: W = P'V'xNap. log. iy]* (2) 

* The Naperian or hyperbolic logarithm of a number, generally written " log^," 
is obtained by multiplying the common logarithm b)'' the constant 2.302585 



56 COMPRESSED AIR PLANT 

The equation may also be written: 

W=PVlog.(|^), (3) 

a form convenient for use in making air compressor calculations. 
When expressed in ft. -lbs. (by putting V in terms of cu.ft., 
and P,P' in lbs. per sq. in.) : 

W=i44PVlog.(|^) (4) 

which is the general equation for the work of compressors oper- 
ating isothermally and without clearance. 

Work of Compressors without Clearance. Adiabatic Com- 
pression. Referring to Fig. 35, the Hne AD represents the 
initial volume, V, of air at normal atmospheric pressure, and 
the hne EC the final volume V, to which the same quantity 
of air is compressed; that is, V^ is the volume after the com- 
pression part of the stroke is completed and before delivery 
begins. In undergoing this change of volume, the pressure 
increases from P to P\ and the resulting compression line AE 
is an adiabatic curve, following the law: 

C 

PV« = PV* = C (constant), or P = r7^ ... (5) 

The total work done in the compressing cylinder is: • 

W=(Wi+W2-W3), (6) 

in which: 

Wi = the work of compression. 

W2 = work required to force the compressed air out of the 
cyHnder, into the receiver. 

W3 = work done by atmospheric pressure on the suction side 
of the piston, while the inlet air is entering the cylinder. 

First.— The work Wi, in ft.-lbs., is: 



Wi=J\44^dV 



THE COMPRESSION OF AIR 



57 



Substituting the value of P. now expressed in lbs. per sq.in., 
from equation (5): 

ncd\ 

Integrating between the limits V and V: 

Wi = i44C[ j:r„ — J (8) 

Dividing the second member of the equation by (—1) and sub- 
stituting for C its value PV": 

144PV" 



Wi = 



^_I LV^(n-l) Y(n-1)_ 



I44PV 



n—1 



Y^(w-i) 



(9) 

(10) 



. P' V" V /P'\^ 
But, smce p- = t^t^ , v^ ^ \ P^ / " ' which, raised to the n—i power, 



gives : 



y(n-l) /p/\w-l 

Y^c^TT) ^ Tp 



Substituting this value in (10) 



Wi = 



144PV 



n— I 



'\ n-\ 



— I 



(11) 



(12) 



Second. — The work W2, of expelling the air from the cylinder, 

= I44PV . (13) 

V P' /V\" 

Multiplying by vr both members of the expression, -p = ( ^^7 ) : 



V'V /\Y-'^ 



But, 



PV ^\y) ' '^^^'^''^^ PV = PV(^ 



V\" 



-1 



r = (p)" and (J 

p/\ n-1 



n-1 



'\ w-1 



hence 



PV =PV( p-) '^ ; which, substituted in equation (13), gives: 

P 



W2 = i44PV = i44PV(%r) « .... (14) 



58 



COMPRESSED AIR PL.\NT 



Third. — The work W3, done by atmospheric pressure on the 
back of the piston, = 144PV (15) 

Taking the algebraic sum of Wi, Wv and W3, from equations 
(12), (14) and (15), and substituting in equation (6): 



W=i44 



PV 



n — i\_ 



T>'\ n-l 

~^ —1 



+PV 



P' 



n- 1 



-PV 



whence, by reducing to a common denominator: 



V\ 



W=i44 
and cancelling: 



7F\^ 



+ (;/-i)PV(^). « -0^-i)PV 



;/ 



W = 



i44PV;z 



LVP 



n-\ 



— I 



(16) 



which is the general expression for the work of single-stage com- 
pressors, with adiabatic compression, and when clearance is zero. 
The relations between the two conditions of compression are 
represented graphically by Fig. 36. By laying off to scale the 







;A9l^ 



Line 
Vacanm 



Fig. 36. 

volumes of air on the horizontal line of the diagram, the corre- 
sponding pressures at different points of the stroke of the piston 
are measured on the verticals. The adiabatic curve rises 
more rapidly than the isothermal, meaning that more work is 
expended. Perfect isothermal compression is unattainable. 
It is only approximated even with the best cooling arrangements, 
and running the compressor at a very slow speed. On the other 
hand, if the air compressed adiabatically could be kept hot 



THE COMPRESSION OF AIR 59 

until used, the loss of the additional work which was expended 
in compressing it would be prevented. But neither can this 
be done. The air is usually conveyed considerable distances 
before it is used, and radiation from the pipes soon reduces the 
pressure to that corresponding with the temperature of the sur- 
rounding atmosphere. In practice, a combination of the two 
theoretical modes of compression is employed, the net result 
depending upon the degree of perfection of the compressor 
and of the cooling arrangements. When compressing in a 
single cylinder to 60 or 80 lbs., and a piston speed not exceeding 
300 ft. per min., it is probable that about one-half of the total 
possible cooling is all that may be expected.* The aim is to 
begin compression with the air at a low initial temperature, 
and to bring the compression line as close as possible to the 
isothermal Hne. The air must be cooled thoroughly during 
compression and before it leaves the cyHnder; any subsequent 
cooling, in the receiver or in the air main, entails loss. 

In ordinary practice the abstraction of heat during com- 
pression is very imperfect. Some distance must be traversed 
by the compressing piston before there is any considerable rise 
in temperature, and until the temperature does rise no cooling 
can be effected. The abstraction of heat does not begin at the 
beginning of the stroke. The temperatures of the intake air 
and of the cooling water are likely to be nearly the same, so that 
all the possible reduction of temperature in any one cylinderful 
of air must take place in a period of time less than that occupied 
in making the stroke. In modern dry compressors of fairly 
large size, and running at full working speed, the compression 
line is usually much nearer the adiabatic than the isothermal 
curve, and often follows the adiabatic curve quite closely. 

The heat produced by compression may be absorbed: 

1. By introducing cold water into the air cyHnder. 

2. By cooling the cyHnder from without, enveloping it in a 
cold-water jacket. 

Machines of the first class are known as '' wet compressors "; 
those of the second, " dry compressors." 

* Frank Richards, " Compressed Air," p. 66. 



60 co:mpressed air flxst 

Values of n in the equation —=(^77). In purely adiabatic 

compression. ;/ = 1.406; in ordinar}* single-cylinder dry com- 
pressors, n is roughly 1.3. while in the best single-stage wet 
compressors i-^ith spray injection"* ;: becomes 1.2 to 1.25. In 
the poorest forms of compressor the value ;/ = i.4 is closely 
approached. For large, well-designed stage compressors and 
efficient intercooling. ;/. referred to the combined indicator 
cards Fig. 39K may be as small as 1.15. 

Work of Two-stage Compressors, without Clearance. The 
air is brought up to a certain pressure in one cylinder; passes 
to an intercooler, in which the temperature of the air is reduced, 
and finally enters a second cylinder, where the compression is 
carried to the desired terminal pressure. The cylinder ratio 
is such that the work is equally di^"ided between the cylinders, 
but changes of conditions of operation other than those con- 
templated may destroy this equahty. In the following dis- 
cussion* it is assumed that the same quantity of work is done 
in both stages. 

An inspection of the diagram. Fig. 37. shows that there must 
be some best intermediate receiver pressure, for which the total 
work of compression -will be a minimum. For, if this receiver 
pressiu*e approach either P or P" (corresponding respectively to 
the points B and G on the compression cun'e\, then would the 
compression approach single-stage work and the entire com- 
pression Une would lie along BCG. But. vdth an intercooling 
receiver at any intermediate point, the broken line BCDE is 
followed, the sa\'ing in work over single-stage compression being 
represented by the area CD EG. 

The net work of the compressor. W. represented by the area 
ABCDEF. is equal to the work of area ABCH of the first 
stage plus the work of area HDEF of the second stage, or W = 
Wi-f-Wo. Let the condition of perfect intercooling be assumed; 

* I desire here to acknowledge the valuable assistance of Dr. Charles E. Lucke, 
Professor of Mechanical Engineering in Columbia University, who kindly gave me 
the use of his original notes on the method of anal>'sis employed in this discussion 
of the theon.- of stage compression and the effect of clearance in the different work 
cycles of adiabatic compression. 



THE COMPRESSION OF AIR 



61 



that is, the hot air discharged from the first cyHnder is cooled 
in the intermediate receiver to the initial temperature of the 
intake air. The work cycle in each cylinder is the same as that 
of single-stage adiabatic compression, as expressed by equation 
(i6), but with two additional symbols for pressures and volumes. 




Fig. 37. — Reference Diagram, Two-Stage Compressor, with no Clearance and 

Perfect Intercooling. 

Let AB = V = initial volume of free air in first cylinder; 
HD= V = initial volume of air in second cylinder; 
OA = P = initial absolute pressure (atmospheric pressure) ; 
OH = P' = terminal absolute pressure in first cylinder, 
assumed to be also the intermediate receiver 
pressure and therefore the initial pressure in 
the second cylinder; 
OF = P" = terminal absolute pressure in second cylinder. 

VYn\ 



Hence: Wi = -^7- (p^l '^ -i . . . first stage 
W2 = — ^];^ — \ p~ / " ~^\' • s^^^^^ stage 



But, assuming the intercooling to be perfect, PV = P'V 



/r 



. (17) 

whence: 

■ (18) 



62 



COMPRESSED AIR PLANT 



Since the best receiver pressure, P', is that for which W is a mini- 
mum, by differentiating and placing the first differential coeffi- 

cient3^v=o- 
ar 






n-1 



— 1 



n-l 



71-1 (PO " n-l (P'O " 



n—i 



n 



n- 1 
p n 



n-l 



« (PO^- 



= 



ra-1 



Whence : 



(P') '^ (P'O " 



n-l 
p n 



(P') 



2n-l 
'\ n 



1 2n— 1 

+ - 



n-l 



or (PO « " =(PP'0 " , from which P' = VpF', an expression 
for the best receiver pressure. 
Dividing both terms by P: 



p/ (PxP'O^ /P 



>//\ \ 



/p/'\H p'' Y" 

But, (-?r) = / — ^ = pT- Substituting these values in equa- 
tion (i8), remembering that P' V'^ = PV and expressing the work 
in ft. -lbs.: 



W = 



2Xi44PV« 



11 — 1 



p^/\ n-l 
P ' 



. . . (19) 



which is the equation for two-stage compressor work, in terms 
of the initial volume and initial and terminal pressures, with 
perfect intercooling and best receiver pressure. 

By a method similar to the above, the expression for the work 
of three-stage compression may also be deduced : 



W 



3Xi44PV« 



n—i 



f'f\ n-l 



in 



(20) 



in which V" is the terminal pressure in the last, or high-pressure, 
cylinder. 

Effect of Clearance in the Compressing Cylinder. In the pre- 
ceding pages expressions are deduced for the work of compression 
with no allowance for the clearance volume of the cylinder. 



THE COMPRESSION OF AIR 



63 



From a mechanical engineering and structural point of view, 
the question of piston clearance is taken up in Chapter VII. 
It is necessary here to discuss the cycles of operation of single- 
stage and two-stage compression with clearance. While the 
work done per unit of air is the same as that shown by the general 
equations for isothermal and adiabatic compression, the work 
per unit of cylinder displacement will be changed, because of 
the re-expansion of the clearance air. In other words, clearance 
affects the volumetric output of the compressor, but not the 
work of compression per unit of volume of air taken into the 
cylinder. 




Fig. 38. 



J F H G 

-Reference Diagram. Compressor Working Isothermally, with Clearance. 



Work of Compressors with Clearance. Isothermal Compres- 
sion. Fig. 38 is a general reference . diagram, in which BC 
represents an isothermal curve. 

LetEB =MF = P; 
GC = FD = P'; 
JE=0B=V; 
CN = V'; 

J F = ND = clearance volume ; 

J H = OA = (OB - AB) =Y'\ or volume occupied by the 
re-expanded clearance air. 

According to the diagram areas: 

Net work ABCD = compression work and delivery work 



OB CN — re-expansion work OADN, 




64 COMPRESSED AIR PLANT 

For the work of a compressor without clearance, and isother- 
mal compression, the general expression, W = PV logJ p- ), has 

already been deduced. This applies to the areas bounded by 
the two horizontal lines, the vertical line and the compression 
line. Similarly, the re-expansion w^ork represented by the area 

/P'^ 

OADN (under the curve DA) =PV''' log, 

Hence, W = PV log/|-) - PV'' 

= p(V-V-)l0ge(^') (21) 

Replacing (V — VO by L, which represents the intake capacity 
of the compressing cylinder, neglecting heating during suction, 
and expressing P in lbs. per sq. in.: 

W=i44PLlog/^) (22) 



Comparing equations (4) and (22) it is seen that they are iden- 
tical, as noted above; but it must be remembered that the volume 
of air actually taken into the cyhnder at each stroke and com- 
pressed is reduced on account of clearance, and hence the volu- 
metric capacity of the compressor is also reduced. Moreover, 
neither in this work cycle, nor in that for adiabatic compression, 
is any account taken of the heating and cooling effects which 
occur during intake and compression, nor of frictional and other 
losses which affect capacity and work per unit of air. These 
points are discussed elsewhere in this chapter and in Chapters 
IV, V, VI, VII and X. 

Single-stage Adiabatic Compression, with Clearance. The 
diagram, Fig. 38, may be used here also, by assuming the line 
BC to be an adiabatic curve. But, though the work areas are 
designated as above, under isothermal compression, and their 
significations are identical, their numerical values are different. 

From Eq. (16), the work corresponding to area OBCN = 

i44PVwr/P'\-"--i 1 
Wi = - — — — ("pj " ~ih and, similarly, the work corre- 



THE COMPRESSION OF AIR 65 



spending to the area OADN = 



W2 = 

Whence, W 



n— I 



n— I 






Replacing (y — V") by L (the intake capacity of the cyHnder, 
with clearance) : 



iaaVIuu 

W = 



n— I 



— I 



(24) 



Since V in equation (16) may be replaced by L, a comparison 
of this equation with (24) shows that the work is the same per 
unit of volume of air admitted to the cylinder; but, the volu- 
metric output is reduced by the clearance. 

Though the pressure-volume formulas serve for most pur- 
poses, it is sometimes convenient to have the work expressed in 
terms of cylinder displacement and volumetric efficiency: 
Referring to Fig. 38: 

Let D = displacement volume of cylinder in cu.-ft., or the 
effective area X stroke, represented on the dia- 
gram by MB =FE; 
C = clearance expressed as a fraction of D ; whence C X 
D = Vc = clearance volume, represented by ND 
= JF, and D (i+C)= total cylinder volume in 
cu.ft., represented by JE; 
V' = volume of re-expanded clearance air ; 
L = intake free air capacity = JE — JH = V — V'; 

E = volumetric efficiency =Y^ = the ratio of the length of 

the actual admission line, AB, to the total distance 
swept through by the piston. 

Then: V = D (i-fC), and V'^' = Vc(^H = Cd(|- K 



Whence V-V''' = D 



i+C 



-c(|)i]=.. 



66 



COMPRESSED AIR PLANT 



Substituting this value of V — V' in equation (23) : 



i44.Vn 



n—i 



i+C-C(^)" 






(25) 



which expresses the work in terms of displacement and clear- 
ance, with pressures in lbs. per sq. in. 



Since E^rp-, L = ED = D 

substituting in (25): 



i + C-C(^V^ 



, as aoove; whence, 



W = -^^— ED 



n—i 



P ' 



(26) 



Work of Stage Compressors, with Clearance. (Fig. 39.) 
The method of analysis of the work of a two-stage compressor, 




Fig. 39. — Reference Diagram. Work of Two-Stage Compressor, with Clearance. 



without clearance (pp. 60-62), applies also here. With clearance 
the work for each stage is expressed by equation (23), the neces- 
sary changes being made in the signification of the different 
factors: 



Wi (first stage j 



i44P(V-V''0 



W2 (second stage) = ^^^ [ [^j " - 1 J J 



. (27) 



tMe compression of air 6? 

In equation (27) the symbols have the same significations as on 
pages 61 and 6;^, in addition to which V^^ is the volume of 
re-expanded clearance air of the high-pressure cylinder. But, if 
intercooling be perfect, 

PV - PV'' = P V - P V^ 

that is, the weight of air entering the second cyHnder is equal to 
that entering the first. Hence, the general equation for the 
work of a two-stage compressor, with clearance and perfect 
intercooling, takes the form: 



^^i_44P(V-V"')» 



n 



/-pf\n-l /p//\ w-1 "j 



By differentiating this equation, and making the first differential 
coefficient t^ = o, it is found that P' = VPP'^ This is the 

value for the best (most economical) intermediate receiver 
pressure. It is the same as that deduced on page 62 for stage 
compression without clearance, since the receiver pressure is a 
function of the compression line and not of the re-expansion line.* 
By substituting in equation (27) the value of P' for best receiver 
pressure, it will be found that, as in the work cycle for stage 
compression without clearance, there is here also an equal 
division of work between the two cylinders. Finally, if the same 
value of P' be substituted in equation (28), this equation takes 
the form: 



^^2Xi44P(V-V'> 



n—1 



(29) 




The diagram (Fig. 39) shows a single, continuous re-expan- 
sion line, FA, which is taken to represent the re-expansion fines 
of both cylinders. This evidently is true only when the clear- 
ances are proportionate ; that is, when the volume of the clearance 
air of the high-pressure cylinder, after re-expansion, is equal to 

* For the sake of brevity, the steps in the deduction of these and the follow- 
ing work formulas are omitted. Readers who desire to pursue the subject further 
will find a full discussion in Lucke's Engineering Thermodynamics . 



68 



COMPRESSED AIR PLANT 



the clearance volume of the intake cyhnder. But the cylinders of 
stage compressors may, and usually do, have different clearances, 
between which no particular relation exists. However, since the 
volume of air dehvered by the low-pressure cylinder must 
necessarily be equal to the volume received by the high-pressure 
cylinder, disproportionate clearance does not affect the work 
done per unit of air compressed. The diagram of the high- 
pressure is merely displaced somewhat with respect to that of 
the low-pressure cyhnder. as shown by Fig. 40, in which FE = 
F'E'andHD=H'b^ 




'\CiDi 

Fig. 40. — Reference Diagram. Work of Two-Stage Compressor, with Dispro- 
portionate Clearance. 

In a manner similar to the above may be deduced the expres- 
sion for work of a tliree-stage compressor, with clearance and 
best intermediate receiver pressures: 



W- 



3Xi44P(V-V"');i 



n— I 



in — 



■ (3°) 



in which P''' is the dehvery pressure, and \'" the re-expansion 
volume of the intake cyhnder. 



THE COMPRESSION OF AIR 69 

The work and capacity of stage compressors may also be 
expressed in terms of displacement and clearance. 

Let Di and D2 = cylinder displacements, respectively of the 

low- and high-pressure cylinders ; 
Ci and C2 = fractional clearances; whence CiXDi and 
C2XD2 = clearance volumes, and Di(i-'^ Ci) and D2(i + C2) = 
total cylinder volumes, all in cu.ft. 

AB 

El =:^:^ = volumetric efhciency of low-pressure cylinder-, 

E2 = TyYy = volumetric efficiency of high-pressure cylinder. 

In the demonstration leading to equation (25) it was found 

'P'\ I 



that V-V''' = D 



i + C-C(p 



Applying the proper sub- 



scripts for two-stage work, and substituting this value of 
V — V' in equation (28), remembering that the volume dis- 
charged by the low-pressure is equal to that received by the 
high-pressure cyhnder : 

w=^'>.[— <l)-l[©-(;)^-4 '3.) 

which expresses the work in terms of displacement, clearance, and 
initial and terminal pressures. 

Since, when the intercooling is perfect, the temperature of 
the air entering the high-pressure cylinder is the same as that 
entering the intake cylinder, it follows that the product of the 
initial pressure in each cylinder and the volume of air admitted 
to each is the same for both cylinders. As previously stated, 
P(V-r'0 = P'(V''-V'^); and, as shown on p. 62, P' = (PP'0^^ 
for best receiver pressure. Therefore, 

P(V- V) = (PP'0^(V''- V"); whence, 



V^'-V"' P VP 



70 COMPRESSED AIR PLANT 

But, since V - V" = DiEi and \" - V^^ = D2E2, 

P'^^ DiEi 



P / D2E2' • • • • • 

According to the reasoning which led to equation (26), 



(32) 



and 



DiEi=Di[i+Ci-Ci(^']«] .... {^^) 

D2E2 = D2[l+C'2-C2(|7)"J .... (34) 



Dividing {^^) by (34) and combining with {^2): 



-" D,E, D,L'+^"'-C'(t)"1 



D2E2 D2! 



(35) 



[l+C2-C2(^)»] 

For convenience in applying this equation, it may take the form: 

^^^^^^ [x+c.-c.(^)4 ■ ■ ■ 

This transformation follows from the relation, for best receiver 
pressure: 

Equation (36) expresses the ratio between the displacements of 
the cylinders, in terms of initial and terminal pressures and 
clearances. 

If the percentage clearances of the cylinders are equal, Ci = C2 
and Ei=E2; whence, the quantities in brackets of the second 



same 



member of equation (36) cancel, and f^ = ( ^ ) • The 

would be true if Ci and €2=0; a condition which may be 
assumed for compressors having very small clearance. 



THE COMPRESSION OF AIR 71 

From equations (35) and (36), for the condition of best 
receiver pressure, several relations may be determined: (i) The 
ratio of compression for a given ratio of cylinder capacities, or 
conversely; (2) The ratio of cylinder displacements for known 
volumetric efficiencies; (3) The ratios of compression in the 
two cylinders which will produce best receiver pressure, the dis- 
placements and clearances being known, or conversely. In the 
third case, several approximations will usually be required. 

For the performance of air compressors see Chap. X. Tables 
are there given, showing the work actually required per cu.ft. 
of free air, for single-, two- and three-stage compression. 



CIL\PTER IV 
WET COMPRESSORS 

Although wet compressors are obsolete in the United States, 
some attention should be given to them, because a few are still 
used in Europe, and a discussion of their design and operation 
will lead to a better understanding of the comparative merits of 
the modes of cooling employed in dry compressors. 

Wet compressors comprise: i. Those in which water is 
introduced in bulk into the air cylinder, and is injected also in 
the form of spray. 2. Those in which water is injected only 
in the form of spray or jets. 

Compressors of the first type comprise some of the earliest 
forms. One of the best is the modernized Dubois-Franfois, 
built at Seraing. Belgium. It has been widely used in Europe, 
for mining and tunnelling, .\nother is the Humboldt Fig. 41). 
A mass of water forming the compressing piston is moved to 
and fro by a plunger C. Connected to each end of the cyhnder 
by an easy curv-e is an air chamber, ha\'ing inlet and discharge 
valves at f and g, made of rubber rings of round cross-section. 
As the piston reciprocates, the air is drawn alternately into one 
air chamber and compressed in the other. At the end of each 
stroke the air compressed by the rising mass of water in the 
chamber is discharged into the receiver. The air is partially 
cooled by contact with the water, and to keep the water cool 
proper circulation must be maintained. Further cooling is 
caused by the injection of sprays into the air chambers from a 
small force pump c. operated from the cross-head d. 

Because of the inertia of the mass of water this t}'pe of 
compressor is generally limited to low piston speeds f 100-150 
ft. per min. or less in some cases). As this is about one-third 
to two-fifths of the piston sp>eed of modern dr\- compressors, 

72 



WET COMPRESSORS 



73 



a wet compressor is heavy and bulky for a given output of air. 
A more recent form of the Humboldt compressor is said to run 
successfully at speeds of 300-360 ft. per min., the temperature 
of the air at discharge being 7.7°-8o° F. These remarkable 
results are open to question for regular, normal service. Lower 
speeds are always advisable for this type of compressor, as 
violent shocks are caused by rurming at high speed. 

The Hanarte compressor is similar in principle to the Hum- 
boldt. Many have been built for French and Belgian mines, 
and also for ice-making plants. They are generally of large 
size, and are efhcient at piston speeds of 250-275 ft. per min. 
The widely splayed out vertical ends of the cylinder cause 




m\ 



— \^ q ^ 

Fig 41. — Humboldt Wet Compressor. 



the water level to rise slowly towards the end of the stroke, 
and afford space in the cylinder heads for large and readily 
accessible inlet and delivery valves. Sprays are also used. 

In wet compressors of this class an efhcient circulation of 
water is difficult to maintain. Only a small quantity of cool 
water can be injected at each stroke, and without copious sprays 
the cooling is imperfect; although the mass of water in the 
cylinder and air chambers is large, there is between it and the air 
only a momentary surface contact. Since water is a poor 
conductor of heat, the air is cooled more by contact with the 
wet cyhnder walls than with the small superficial area of the 
rising and falHng water. Furthermore, the compressed air is 
practically saturated with moisture. 

Compressors of the second type, in which cooling water is 
used only in jets or spray, are much less cumbrous than the older 



74 COMPRESSED AIR PLANT 

design and permit a higher piston speed. They were first built 
by Golladon, at the St. Gothard tunnel. Though some of them 
are still used in Europe, they are obsolete in American practice. 
The air cyhnder does not differ materially from that of the dry 
compressor. A water pipe is tapped into each cylinder head 
and fine spray is injected in front of the piston during com- 
pression. Since the water is in a state of fine division a relatively 
large surface contact is presented, and the air is thoroughly 
saturated with moisture during compression. Zahner states 
that CoUadon's St. Gothard compressors, " which were run at a 
piston speed of 345 ft. and compressed the air to an absolute 
tension of 8 atmospheres (103 lbs. gage pressure), gave an 
efficiency which never descended below 80%, while the tem- 
perature of the air never rose higher than i2°-i5° C. (53°-59° 
F.)." The temperature of the injection water is not 
stated, but must have been very low to obtain these 
results. 

A dry compressor may be converted into a wet compressor 
merely by providing the water jets. The injected water col- 
lects in the cylinder until the clearance space at the end of the 
stroke is filled. The surplus is forced out at each stroke with 
the compressed air through the discharge valves, and is drained 
away from the receiver. As the piston clearance in well-designed 
compressors is very small, little water remains in the cylinder 
to be churned back and forth by the piston. The injection water 
should be pure and as cold as possible. Gritty water injures 
the cylinder, piston and valves. 

In proper injection apparatus: i. The injection must com- 
mence at the beginning of the stroke and continue to the end, 
against the advancing piston. 2. There should be thorough 
diffusion of the spray throughout the cyhnder. By mere surface 
contact water takes up but little heat. Even a single strong jet 
is quite effectual, because on striking the piston it is broken into 
spray. 3. The volume of injected water should increase with 
the air pressure produced, that is with the quantity of heat 
generated. With insufficient water much moisture is taken 
up by the warm air and carried into the receiver. 



WET COMPRESSORS 



75 



The quantities of water required for different pressures are 
shown in Table III.* 



Table III 



Pressures. 


Heat units Generated 
by Compression in 
I Lb. of Free Air. 


Pounds of Water to be Injected a\ 

68° F. to Keep Final Temperature 

at 104° F. 


Above 

Vacuum, 

Atmospheres. 


Gage 

Pressure, 

Lbs. 


Per Lb. of Free 
Air. 


Per Cu.ft. of Free 
Air. 


2 


14-7 


58.310 


0.734 


0.056 


3 


29.4 


92.390 


1. 164 


0.089 


4 


44.1 


116.627 


1.469 


112 


5 


58.8 


135-388 


1 . 701 


0.130 


6 


73-5 


151.700 


1. 891 


0.144 


7 


88.2 


163.735 


2.063 


0.158 


8 


102.9 


174-937 


2. 204 


0.168 


9 


117. 6 


184.865 


2.329 


0.178 


lO 


132.3 


193.701 


2.440 


0.186 


12 


161. 7 


209 . 090 


2.634 


0. 201 



* This table is taken in part from that given by Zahner, " Transmission of 
Power by Compressed Air," p. no, English units being substituted for French. 



CHAPTER V 

DRY COMPRESSORS 

In dry compression no water enters the air cylinder except 
that which is carried as moisture in the air itself. All the 
cooling, aside from radiation, is effected by water-jacketing 
the cylinders. 

Water-jackets. Fig. 42 shows the longitudinal section of a 
Nordberg cylinder. (Eigs. 4, 8, 24, 26, 27, and other cuts of 
longitudinal sections, illustrate different types of jacketed 
cylinders.) The annular space JJ is occupied by water, and 
nearly one-half the area of each cyhnder head is also covered 
by water jackets KK. The remainder of the end areas is 
occupied by the suction and delivery valves. Circulation of 
water is effected by pipes connecting with the openings A and B, 
respectively for inlet and discharge. To assist circulation the 
jacket spaces are subdivided. Cold water enters at A, and, after 
passing through the annular and end jackets JJ, KK, is dis- 
charged at B. For maximum cooling effect, the jackets on the 
cylinder heads surround the valves and air passages as completely 
as possible. C is a drain pipe for blowing out sediment. 

In some designs, the annular jacket is divided by vertical 
partitions, so that the cold water entering at the top passes 
first around about one-fifth of the length of the cylinder nearest 
each end; then around the middle portion, and is discharged 
at the top. This arrangement recognizes that at the end of the 
stroke, where the air pressure is highest, the greatest amount 
of heat is generated. In other designs but little of the cyHnder- 
head area can be jacketed, because of the space occupied by 
the inlet and discharge valves. This would seem to be a defect 
because, on approaching the end of the stroke, the piston rapidly 

76 



DRY COMPRESSORS 



77 



covers the annular jacket, leaving only a small part of its area 
available for cooling the hot air during discharge. It is at 
this point of the stroke that large end jackets are most valuable. 
In still other compressors, the water passes first into the end 
jackets, and then successively through separate compartments 
of the annular jacket. The deHvery valves of some compres- 
sors are placed radially, close to the cylinder ends, whereby a 




Fig. 42.— Air Cylinder of Nordberg Compressor. 



larger proportion of the heads can be jacketed (Fig. 43). The 
jacket in Fig. 44 has 8 longitudinal partitions, extending alter- 
nately from each end of the cylinder nearly to the opposite end. 
The water, which enters near the top, is forced to travel back 
and forth between the partitions and from end to end of the 
cylinder until it is finally discharged. The cooHng watet is 
often taken from a tank, set at an elevation above the com- 
pressor, or a small pump may be employed. 

The useful effect of water-jackets depends largely on the 
running speed of the compressor. In the best single-stage 
compression, to say 70 or 75 lbs. and at not over 300 ft. piston 
speed, probably not more than about one-half of the total 
possible cooKng can be effected; that is, n would be equal to, 
say, 1.25. Heat is generated faster than it can be abstracted, 



?8 



COMPRESSED AIR PLANt 



since only a part of the air passing through the cylinder corned 
into direct contact . with the cooKng surfaces. The cylinder, 
discharge pipe, and even the receiver, are usually quite hot 
when the compressor is running at full speed; often too hot 
to be touched with the hand. With well-jacketed cyHnders, 
and compressing only to 45 lbs., the temperature of the air at 
delivery has reached 280° F. The heat of compression in 
dry compressors probably ranges from 20o°-40o° F. for the 



.S;39' *ff^datiBiftit:MfA^ i 




Fig 43. — Air Cylinder of Allis- Chalmers Compressor, with Nine Delivery Valves 

Set Radially at Each End. 

ordinary pressures used in mining, though not often exceeding 
350°. The temperature should not be allowed to rise above this 
point.* At a mine in Montana, the writer has observed the thin 
wrought-iron delivery pipe of a 50-drill compressor red-hot 
for a distance of nearly 6 in. from the cylinder shell. Driving 
compressors at too high a speed often causes the poor results 
complained of by some users of compressed air. 

* T. G. Lees, Trans. Federated Inst. Mining Engrs., \'ol. XIV, p. 569. See 
also Chapter XIII of present volume. 



DRY COMPRESSORS 



79 



Jacket 



Suction 



The inner shell of the air cylinder, i.e., between the cylinder 
and jacket, has sometimes been made of hard brass, which by 
its high conductivity assists in carrying off the heat. With the 
same end in view, the cylinder walls should be as thin as is 
consistent with safety. Be- 
sides cooling the air during 
compression, the water-jacket 
of a dry compressor is indis- 
pensable in keeping down the 
temperature of the cylinder 
shell. Without jackets the 
metal of the cyhnder would 
become hot enough to burn 
the oil, and render proper 
lubrication impossible. 

Piston Clearance in the 
Air Cylinder. Toward the 
end of the stroke, the com- 
pressed air in front of the 
piston begins to pass out 
through the dehvery valves 

as soon as its tension exceeds pj^, ^^ _Air Cylinder, Class " E/' Laid- 
that of the air in the dis- law-Dunn-Gordon Compressor. 

charge pipe to the receiver. 

But, in a dry compressor on completion of the stroke, a certain 
quantity of hot compressed air remains in the clearance space. 
On the back stroke this clearance air expands behind the piston, 
and no fresh air can enter through the inlet valves until the 
cylinder pressure falls below atmospheric pressure. Hence, 
it is never possible, in a dry compressor, to take a full cylinder 
of fresh air at each stroke; that is, the volumetric capacity per 
stroke, in terms of cu.ft. of free air, is always less than the 
volume swept through by the piston. In a wet compressor 
the clearance space is filled with water, and does not effect 
the volumetric capacity. 

Fig. 45 shows the effect of clearance. Before the inlet 
valves can open, the piston must travel from c to b, and the cor- 




80 



COMPRESSED AIR PLANT 



responding cylinder volume passed through by the piston repre- 
sents the loss of volumetric capacity. The actual effect of 
clearance on the volumetric efficiency of the compressor depends 
on the dehvery pressure. The higher this pressure, the greater 
is the distance cb. The loss of volumetric capacity, although 
important in the operation of the compressor, does not involve 
a corresponding loss of useful work (see under Eq. 24, Chap. III). 
The compressed air remaining in the clearance space helps to 
overcome the inertia of the moving parts at the beginning of the 
return stroke, and to compress the air on the other side of the 
piston. The clearance air cools slightly during the momentary 
stoppage of the piston as the stroke is reversed, but the conse- 




Air Card showing 
effect of clearante. 
Volume between b and c 






-j-Atmospheric ' 
l) Ic Line * 

Vacuiuu. ■ 



Fig. 45 



quent reduction of pressure is neghgible. In expanding behind 
the retreating piston, however, the clearance air cools rapidly 
and does not tend materially to raise the temperature of the 
incoming atmospheric air. 

The effect of clearance in reducing the capacity of a dry 
compressor is shown by Fig. 46. For clearances greater than 
1% the loss is serious, even at pressures of 75-100 lbs. 

In cylinders of the same diameter and having the same amount 
of linear clearance, the ratio between cylinder volume and 
clearance volume depends on the length of stroke. This ratio 
is generally largest in short-stroke compressors and smallest in 
those of long stroke. It varies, also, in compressors of different 
makers. Several Ingersoll-Rand compressors have the follow- 
ing ratios between cylinder and clearance volumes: 



DRY COMPRESSORS 



81 



14-in. stroke 0190 36-in. stroke. ........ .0112 

21 " " . .0176 48 ^' " 0093 

24 " " .0126 

ranging thus from about 2% down to 1%. Some compressors 
of the same makers, of 42-in. stroke, but relatively small cylinder 
diameters, have clearances as small as .78, .80 and .90 of 1%, and 
several of 36-in. stroke have clearances of .S^ and .84 of 1%. 



250 



225 



200 



175 



150 



^ 100 

cs 

O 

75 



50 



25 





/ 




/ 




/ 


/ 




/ 




( 




/ 


/ 


/ 




/ 






/ 






0/ 


a/ 




/ 




/ 






/ 






£1 

? 




" 1 c 




f 


/ 


/ 










^/n? 






r ^/ 


If 


V 














/ 


/ 

/ / 




















// 


V/ 


7 














' 




// 


// 




















r 











































10 15 20 25 30 35 40 45 
Percentage of Piston Displacement 



50 



Fig,' 46. — Diagram Showing Effect of Piston Clearance {Eng. News) 



In some Laidlaw-Dunn-Gordon compressors clearances range 
from .75 to 1.25%. Clearances are generally larger; thus in the 
direct, electric-driven, two-stage, Ingersoll-Rand compressors, 
the following clearances are found in the intake cylinders: 

28X24 in . 2.15% 18X14 in 2.00% 

23X20'' 1.70 17X14" 2.19 

19X16'' 1.85 

Compressors of some other makes have clearances as large as 
2.5%. With few exceptions the lowest figures apply to large, 



82 



COMPRESSED AIR PLANT 




Fig. 47. 



long-stroke compressors; the higher to the small, short-stroke 
machines. 

Fig. 47 indicates the method of minimizing clearance for 
ordinary pistons, by casting a recess in the cyhnder head to 
receive the piston-rod nut at end of stroke. With small clear- 
ances the compressor must have 
careful attention, so that, if the 
working length of the connecting 
rod should be varied in fitting 
new brasses, the piston will not 
strike the cylinder head. 

Examples of other devices for 
overcoming the disadvantages of 
piston clearance: 

I . Longitudinal bye-pass 
grooves (Fig. 47, B) are cast 
in the inner surface of the cylin- 
der near the ends, so that when 
the piston reaches the end of 
its stroke the grooves are partly 
uncovered, and the clearance air passes to the other side of the 
piston. 

2. In sHde- valve compressors the valve may be provided with 
a " trick-passage," which at the end of the stroke is brought 
into connection with two small ports entering the extreme 
ends of the cyhnder, thus releasing the high-pressure clearance 
air into the other end of the cyhnder. 

An objection to releasing all the clearance air is that the 
sudden removal of the heavy pressure on the piston causes 
hurtful shocks. In recent American compressors the clearance 
space is very small, and the air confined in it is not 
released. 

Dry versus Wet Compression. By bringing the air into 
direct contact with water the heat is most effectually absorbed, 
provided the injected water is properly applied (Chap. IV). 
Without cooling, the work converted into heat during corn- 
pression, and therefore lost, is as follows; 



DRY COMPRESSORS 



S3 



Compression to 2 atmospheres, 9.2% loss, 



" ■ 3 


15.0% '* 


4 


19.6% " 


5 


21.3% " 


6 ' 


24.0% - 


7 


26.0% *' 


8 


27.4% '* 



In well-designed dry compressors, working at 5 atmos- 
pheres, the heat loss is reduced about one-half (from 21.3% 
to 11%), while in ordinary mining practice, with single-stage 
compressors, the loss is often fully 15%. By spray injection this 
loss has been cut down to as little as 3.6%, and in some large, 
slow-running European wet compressors to 1.6%. But, low first 
cost and simpKcity of construction may be more advantageous 
than a close approximation to isothermal compression. There 
are two considerations: (i) the effect of injected water upon the 
compressed air and the machines using it. (2) the effect of the 
water upon the working of the compressor. 

First, by using large slow-speed engines, and an abundance 
of injection water, the air is well cooled, though at a higher 
first cost for plant. Wet compression gives a good indicator 
card. Table IV shows that in compressing moist air somewhat 
less work is expended than for dry air. This is because the speci- 
fic heat of watery vapor is about twice that of dry air; therefore 
in the presence of moisture more heat is required to raise the 
temperature of the air in the compressing cylinder. 



Table IV 



Absolute Pressure, 
Atmospheres. 


Gage Pressure, 
Lbs. 


Ft. -Lbs. of Work to Compress i Lb. Air. 


Dry Compression. 


With Sufficient 
Moisture. 


I 
2 

3 
4 
5 
6 

7 



14.7 
29.4 
44.1 
58.8 

73-5 
88.2 


23,500 
37,000 
48,500 
58,500 
67,000 
75,000 


22,500 
35,000 
45,000 
52,500 
60,000 
66,000 



84 co:mpressed air plaxt 

Theoretically, a corresponding economy takes place when 
the air is expanded again in the machine using it. 

Objections to Wet Compressors. The amount of heat 
absorbed during compression is proportional to the difference 
of temperature between the intake air and the injected water, 
and to the time of contact between the air and water. This 
difference of temperature is usually zero at the beginning of 
the stroke, reaching its maximum at the end. Hence: (i) 
to attain a fair approach to isothermal compression the piston 
speed must be very slow; (2) during the first part of the stroke 
but Httle heat is removed, and it is only when compression is 
complete, and discharge from the cyhnder begins, that the 
cooling effect is at its maximum. At ordinary piston speeds, 
therefore, a large proportion of the total heat must be given 
up after the discharge valves have opened; in other words, 
after compression is completed. So far as economy of work is 
concerned, the lower hnal temperature due to spray injection 
is in a measure deceptive. The warmth of the air at discharge 
augments its moisture-carrWng capacity, and the separation of 
the water in the receiver is of necessity imperfect in a receiver 
of any reasonable size. ^luch moisture passes into the air 
mains, deposits as the air cools in long pipe lines, and in cold 
weather may freeze so as to reduce the eft'ective diameter of 
the pipe. Moisture remaining in the air has a further ill eft'ect 
when it is used. At the instant of exhaust by the drill, or other 
air engine, the cold produced by expansion may cause trouble- 
some accumulations of ice in the exhaust passages. 

In the dry compressor, since air is a poor conductor of heat 
it can give up but httle of its heat of compression between the 
piston strokes. But, although the moisture always present in 
atmospheric air will make its appearance as frost at the exhaust 
of the air machine, there is rarely enough of it to cause serious 
trouble.* The delivery of warm air by a dry compressor 

* The quantity of moisture in the atmosphere, or its humidity, varies with the 
climate, the season of the year, and in a measure with the altitude above sea-level. 
It is usually greatest near the ocean or any large body of water. WTiat is commonly 
called dr>- atmospheric air contains from 40 to 50*^ of the quantity necessar>' to 
saturate it. The degree of saturation in summer often reaches 90*^ or more. 



DRY COMPRESSORS 85 

is far less objectionable than warm air from a wet com- 
pressor. 

Second, as to the effect of injected water upon the working of 
the compressor. Water in the air cyHnder is always objection- 
able, because it makes lubrication difficult, causes rust, and 
increases the wear of piston and cylinder. There is no satis- 
factory method of lubricating wet compressor cylinders. 
Injected water must be pure and free from grit. Water that is 
harmless for use in jackets might be injurious to valves, cylinder 
and piston. Mr. W. L. Saunders states that, although the 
thermal loss is higher in dry than in wet compressors, the friction 
loss is considerably higher in the wet compressor. The net 
economy of the best wet compressors is probably no greater than 
that of the best dry compressors. 



CHAPTER \1 
COMPOUND OR STAGE COMPRESSORS 

Compound or stage compressors divide the work of com- 
pression between two or more cylinders. In two-stage com- 
pressors air at atmospheric pressure is taken into the large or 
low-pressure cylinder, is there compressed to a certain point, and 
is then forced into the high-pressure cylinder, where it is brought 
up to the required tension. The cylinders are proportioned 
to di\'ide the total work equally between them. This secures 
equahzation of resistances, and promotes the efficiency of the 
cooling apparatus. The theory is given in Chap. ITT. Since 
the heat of compression increases \^'ith the pressure produced — 
though not proportionately — it becomes difficult at high pressures 
to keep do\Mi the temperature to a point permitting proper 
lubrication of the air cylinder. In attempting to compress 
even to 90 lbs. gage in one cylinder, the theoretical final tem- 
perature becomes 459° F. Though some heat is dissipated by 
radiation, the working temperature corresponding to this pres- 
sure may still be too high to be dealt with effectually by the 
ordinary water-jacket, because in a single cylinder the area to 
which cooKng can be applied is too small relatively to the volume 
of air, and the total compression period too short. Even when 
working at moderate piston speed (350-400 ft. per min.), the 
cooling is so imperfect that the compressed air at dehvery is 
ver\' hot, causing considerable loss of pressure and work due to 
subsequent cooling. 

Formerly, stage compression was employed for high pressures 
only, as for pneumatic locomotives, riveting machines, presses, 
compression of gases, etc. To produce vev}' high pressures (500- 
1,000 lbs. or more) three- and four-stage compressors are 
necessary. 

86 



COMPOUND OR STAGE COMPRESSORS 87 

It is now recognized that two-stage compression is advan- 
tageous even for pressures of 70-80 lbs., as commonly used for 
machine drills. The cooling during compression is more 
thorough, because the total heat generated is divided between 
the cylinders; in each the temperature is lower than when the 
same total pressure is produced in a single cylinder, and the 
combined water-jackets afford a larger cooling surface. 

A further cooling is effected by an " inter cooler " (Figs. 
49, 50), placed between the cylinders. It is an intermediate 
chamber, through which the air from the intake or low-pressure 
cylinder passes on its way to the high-pressure cyhnder. The 
temperature of the air is here reduced, so that when the high- 
pressure piston begins its work the temperature of the volume 
of air on which it acts is considerably below that at which the air 
was discharged from the low-pressure cylinder. The total 
reduction of temperature depends on the volume of air under 
compression, the area of cooling surfaces and the length of time 
the air is in contact with these surfaces, which factor in turn 
depends on the piston speed. 

Range of Work for which stage compression is applicable: 

I'. Although stage compression is theoretically advantageous 
for all pressures, it is of doubtful utihty for gage pressures of 
much less than 75 lbs., because of the small saving as compared 
with the greater cost of the more complicated mechanism. 
It is generally applicable for pressures higher than 70-75 lbs. 

2. Stage compression is especially useful for large plants, in 
which the percentage of saving will represent an amount suffi- 
cient to warrant the greater first cost. 

3. The higher thermodynamic efficiency of stage compression 
is partly offset, and in poorly designed plants may be entirely 
neutralized, by the increased frictional losses due to the use of 
several cylinders. A well-designed, economical steam end should 
be used, together with efficient cooling for the air end; other- 
wise stage compression may cost more per cu.ft. of air deliv- 
ered than simple compression in a well-designed compressor. 

All ordinary stage compressors are double-acting; that is, 
on each forward and back stroke air is taken into the cyUnders 



88 CO]MPRESS£D AIR PLANT 

on one side of the piston, while compression and dehvery are 
going on on the other side. In the single-acting form the 
resistances in the cyHnders are not as well equalized throughout 
the stroke. It is employed for some kinds of service, as for the 
high-pressure cyHnders of locomotive chargers (Chap. XXVI), 
largely because of the difficulty of maintaining air valves for 
very high pressures; also to simplify the heavy castings re- 
quired, and to give the water jackets more opportunity to act. 

Double-Acting Two-Stage Compressors. This type is more 
satisfactory than the single-acting compressor, because the 
cycle of operations during each forward and back stroke is the 
same, and the distribution of resistances is more uniform. 
Three forms may be taken to represent accepted practice, viz.: 
the straight-line, two-stage compressor (Figs. 5-8) and the du- 
plex forms, consisting of a pair of staged air cylinders, placed 
tandem to twin-simple, or cross-compound steam cylinders 
(Figs. 9-14). The last-named is best for large plants. 

Fig. 48 shows diagrammatically a two-stage, straight-line 
compressor. Assuming that the pistons have reached the end 
of their forward stroke, the conditions in the two cyHnders are 
approximately as follows: The intake cylinder (D) is fall of air, 
practically at atmospheric pressure, while the high-pressure 
cyHnder (G), with the intercooler (F) and connecting passages, 
are occupied by air just delivered from the intake cylinder, at, 
say, 30-35 lbs., or somewhat less than one-half the final pressure. 
On the reverse stroke the free air in front of the intake piston 
is compressed to 30 lbs. and delivered into the intercooler and 
high-pressure cyHnder, while the air already occupying the lat- 
ter is brought to the final pressure and discharged. 

In standard tandem, two-stage compressors, the volumetric 
capacities of the low- and high-pressure cyHnders are to each 
other in the ratio of about 10-4, the intention being to pro- 
portion them so that their ratios of compression are nearly equal. 
Thus the distribution of work and the heat generated in the 
cylinders will be equalized and most effectually dealt with by 
the intercooler. Practice as regards the relative volume of the 
intercooler and cylinders has changed greatly in recent years. 



COMPOUND OR STAGE COMPRESSORS 



89 




(L) OJ Qj (U 






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90 COMPRESSED AIR PL.AXT 

In recognition of the importance of thorough intercooling, and 
the fact that the first cost of even a very large intercooler 
is moderate, while its maintenance is practically nil, it is now 
made of much greater capacity than formerly. The hot air 
from the intake cylinder is kept longer in contact \\ith the 
cooling surfaces, because of its reduced speed of flow through 
the larger cross-sectional area of the intercooler, and it enters 
the high-pressure cylinder at a correspondingly lower tempera- 
ture. The connections between cyhnders and intercooler 
should be of as small volume as is consistent w^th reasonable 
frictional resistance to the flow of air through them; because 
the air occup}'ing these passages at any given time is exposed to 
but little cooling save that due to radiation. 

It may be assumed in good practice that, if the volume 
of the intake cylinder be lo, then the volume of its connection 
w^th the intercooler should be, say, 1.5, of the intercooler 4, 
of the connection to the high-pressure cylinder 1.5, and of the 
high-pressure cylinder 4 (the net capacity of the intercooler 
may be even greater than is here assumed). Having these 
proportionate volumes, the follo\\ing cycle of operations \^^ll 
take place during a single stroke. Suppose this stroke to be 
from right to left, as indicated by the arrows in Fig. 48. 
By the pre\dous stroke (left to right) the intercooler and its 
connections to the cylinders, representing a volume = 1.5 +4+ 1.5, 
were filled \\ith air compressed, at. say 30 lbs. This body of 
air was then shut off from both cylinders by their valves, and 
lost part of its heat and pressure by the action of the inter- 
cooler. During the first part of the following (left-hand) 
stroke, the intake piston acts only on the cylinderful of free 
air just taken in (volume = 10).* While this is being com- 
pressed, the advance of the high-pressure piston causes the 
air in the intercooler and its connections to begin to flow into 
the high-pressure cyhnder, thereby increasing in volume and 
decreasing in pressure, until a point a Httle beyond mid-stroke 

* The method of analysis here given is s'jnilar to that employed by Frank 
Richards, '* Compressed Air," pp. 86-87, though the quantities used are taken 
to represent a closer approach to current practice in the proportions of the parts. 



COMPOUND OR STAGE COMPRESSORS 91 

is reached. Beyond this point the pressure in front of the 

intake piston rises sHghtly higher than that in the intercooler 

and the dehvery valves open, so that the piston acts upon the 

. lo 4 

entire body of air: volume = — hi.5+4+i.5H — =14. Then, 

2 2 

until the end of the stroke, both cylinders are in communica- 
tion through the intercooler, i.e., from the left-hand end of the 
intake to the right-hand end of the high-pressure cylinder, and 
an approximate equalization of pressure is established through- 
out. 

Until the left-hand, intake delivery valves open, the air in 
the intercooler is isolated from the intake cylinder, in which 
compression has progressed without other cooling than that of 
the cylinder jacket. But when the warm air begins to pass 
through the intercooler into the high-pressure cylinder, the 
influence of the intercooler is exerted upon a new body of air. 
At the end of the left-hand stroke the closing of the delivery 
valves again shuts off the air in the intercooler from both cylin- 
ders. The high-pressure cylinder, on the right-hand side of the 
piston, is occupied by a body of air the temperature and pres- 
sure of which have been reduced by the combined effect of 
the intercooler and both water-jackets to a point below that 
due to the working pressure of the intake cylinder. 

In the latter part of the left-hand stroke, when the intake de- 
livery valves have opened and the piston of this cylinder is act- 
ing on the volume 14, as stated above, part of this air (volume = 

2~}~I.'> 

— = 25% of the total) has passed beyond the influence 

14 

of the intercooler, and another part (volume = ^==46%) 

14 

has not yet reached it. At the end of the left-hand stroke 
the volume of air in the intake cylinder = 0, in the intercooler 
and its connections i.5+4-fi.5 = 7, and in the high-pressure 
cylinder 4, a total of 11, of which 1.5 has not reached the inter- 
cooler, but has been affected only by the water-jacket of the 
intake cylinder. 

This analysis emphasizes the importance not only of employ- 



92 COMPRESSED AIR PLANT 

ing a sufficiently large intercooler, but also of making the con- 
necting passages small. The useful effect of the small inter- 
coolers often used on straight-line compressors should not be 
exaggerated. The best economy is obtained only by cooling 
during compression and before the air leaves the cylinder. 
One-half of the total work of compression — that performed in 
the high-pressure cylinder — is done solely under such cooling 
influence as is afforded by the water-jackets of this cylinder. 
The jackets of both cylinders should, therefore, be of large area, 
with an efficient circulation of cold water. They should cover 
not merely the cylinder barrels, but as much of the heads as 
the space occupied by the valves permits. In the latter respect 
some compressor designs are deficient. 

The details of the distribution of the air in the foregoing 
description apply exactly only to compressors in which the air 
cylinders are tandem to each other. In duplex stage-com- 
pressors, the cycle of operations is different because the pistons, 
instead of moving together in the same direction, work with 
one crank 90° in advance of the other. 

Though the total work done should be equally divided between 
the air cyHnders, still, by reason of the frequent variations in 
receiver pressure, upon which depends the actual terminal 
pressure of the high-pressure cylinder, an approximate equaliza- 
tion only is attainable. On the basis of some terminal pressure 
taken as normal, such diameters are assigned to the cylinders 
as will make their compression ratios equal, or nearly so. Take, 
for example, a pair of cyhnders, 15 ins. and 24 ins. diameter, to 
produce 85 lbs. gage pressure. Assuming that the air between 
the stages is cooled to the original temperature, the absolute 
intake pressures of the cylinders will be inversely proportional 
to the squares of their diameters, or: 15- : 24^ : : 14.7 : 37.64. 
The absolute pressure of 37.64 lbs., as dehvercd by the intake 
cylinder, theoretically equals the intake pressure of the high- 
pressure cylinder. The ratio of compression in the intake 

cylinder is — ^— = 0.3905; in the high -pressure cy Under, — '- — = 

37-64 ^ ^ ^ 99-7 

0.3775. This is as close to perfect equahzation as is necessary. 



COMPOUND OR STAGE COMPRESSORS 93 

The Intercooler in its usual form is a long cylindrical cham- 
ber, containing parallel, thin brass or wrought-iron tubes, 
through which cold water is circulated. The air passes through 
the spaces between the tubes. The intercooler is placed in a 
convenient position between and usually above the cylinders 
(Figs. 8, 23), and as close to them as possible, so that the con- 
necting passages may be short and of small volume, because, 
as already stated, the air in these passages at any given time is 
denied the cooling effect both of the cylinder jackets and of the 
intercooler itself. The intercooler tubes must be close enough 
together thoroughly to split up the body of air traversing the 
intermediate spaces and so secure the maximum cooling effect. 
It is intended that the temperature of the air, on passing from 
the intercooler to the high-pressure cylinder, shall be reduced 
nearly to the normal. The effect upon the compression curve 
of this drop in temperature is shown by Fig. 5 1 ; the high-pres- 
sure compression curve should, and often does, begin close to the 
isothermal line. Every 10% decrease in the temperature of the 
air delivered to the high-pressure cylinder decreases by about 
1% the power required for compression. Thorough cooling has 
therefore been sought by increasing the volume and efficiency 
of the intercooler. 

Brass intercooler tubes are preferable to iron because of 
their higher conductivity; but iron tubes cost less, and being 
rougher present a larger cooling surface to the air flowing between 
them. They should always be as thin as is consistent with the 
necessary strength. The tubes are expanded into tube-plates 
at each end, and by baffle-plates the air is compelled to pass 
through the entire volume of the intercooler. One tube-plate 
is attached to the shell, the other being free to move as the tubes 
expand or contract. The end water-heads are so divided that 
the water circulates actively back and forth several times, 
before final discharge. Fig. 49 shows a recent design. Air from 
the low-pressure cylinder entering at A passes alternately above 
and below the successive bafde-plates to the connection at C 
with the high-pressure • cylinder. To compensate the decrease 
in air volume due to cooling, the baffles are spaced closer towards 



94 



COMPRESSED AIR PL.'VNT 



the high-pressure cyHnder; this maintains active circulation 
between the tubes and prevents excessive pressure drop. The 




■i-t 



U 

a 
a 

<u 
C/3 



CO 

<u 

o 

a> 
r: 



4> 

to 
a 






o 



pressure drop allowed in this design corresponds to 3 J or 4 ins. 
of water column (=0.126-0.145 lbs. per sq. in.). The air may 



COMPOUND OR STAGE COMPRESSION 95 

be cooled to within, say, iS"^ F. of the cooling water temperature 
by using 2| gals, of water per loo cu.ft. of air, or within 15° F. 
by 4 gals, per 100 cu.ft.* The baffle at M forms a trap for 
collecting the moisture deposited due to coohng, the water being 
drawn off by a drain cock. 

The following comparison of the work done by single- and 
double-stage compressors shows the results of thorough cooHng. 
Frictional losses are omitted in each case, and no account is 
taken of the coohng due to the cylinder water-jackets. 

1. A single-stage compressor, producing a gage pressure of 
70 lbs. at sea-level, with a 24-in. cylinder and a piston speed 
of 400 ft. per min., will have a capacity in terms of free air at 
normal temperature of 1,256 cu. ft. per min. For adiabatic 
compression, the mean cyHnder pressure will be SS-^S lbs. and 
the H.P. 184.38. 

2. For doing the same work in a two-stage compressor, 
having an intercooler capable of reducing the temperature of 
the air to the normal between the cylinders, it may be assumed 
that the intake cylinder has the same diameter^ 24 ins., and 
that the pressure produced in it is 35 lbs. The mean pres- 
sure (adiabatic) corresponding to 35 lbs. terminal pressure is 
25.6 lbs., and the H.P., 118. 19. The diameter of the high- 
pressure cylinder, under the assumed conditions, is found by 
making the piston area inversely proportional to the increase in 
absolute pressure of the air delivered to it by the intake cyHnder, 
i.e., in the ratio of 14.7 : 35 + 14.7 = i : 3.38. This gives an area 
of 135 sq. ins., or 13 ins. diameter. Compressing in this cylinder 
from 35 to 70 lbs. gage, the mean effective pressure will be 
28.74 lbs., and the H.P., 46; or a total for both cyhnders of 
ii8.i9-f46= 164.19 H.P. 

Compared with the power required for the same work in 
a single cyHnder, this shows a saving of: 184.38 — 164.19 = 
20.19 H.P., or about 11%. The theoretically perfect cooling 
between the cylinders here assumed is unattainable, and the 
frictional loss in the stage compressor would probably be a little 
greater than in the single- cylinder machine; so that the net 

* Charles A. Hirschberg, communication to the autho?. 



90 



CO:\IPRESSED AIR PLANT 



gain due to intercooling may in this case be taken at, say, 7-8%. 
The saving is increased in deahng with higher pressures. (For 
'' Stage Compression at High Altitudes," see Chap. XIII.) 

Fig. 25, Chap. II, shows in some detail a Sullivan 3-pass 
intercooler. Fig. 50 shows an Ingcrsoll-Rand Vv.tical inter- 




I'lG. 50. — Vertical Intercooler. Ingsrsoll-Ran'^- Cc. 



cooler, with all of its pipe connections. These coolers may be 
used also as '' receiver-after-coolers, " now considered as essential 
adjuncts of well-installed large plants (see end of Chap. XI) 
A similar appliance may be employed advantageously as an 
ante-cooler for the intake air. ' 

Air Cards of Two-Stage Compressor. The compression 



curves of a 
the adiabatic 



two-stage 



compressor are shown in Fig. 



and isothermal curves 



being 



also 



51, 
laid 



COMPOUND OR STAGE COMPRESSORS 



97 



down.* These cards (not accurately reproduced here) were 
taken from a pair of yi and i4Xi6-m. cyHnders, compressing 
to no lbs. gage, at 135 rev. per min., or 360 ft. piston speed. 
Initial temperature of cooHng water, 55°; temperature at dis- 
charge from jackets and intercooler, 62° F. Several points 
are to be noted in connection with these cards: 

First. The overlapping of the high- and low-pressure cards 
indicates a loss, because the work represented by the area of 
overlap is in reality work done twice. This is the result of the 
drop in pressure between the cylinders, caused by the resistance 




Line 



Fig. 51. — Combined Air Cards of Two-Stage Compressor. 



presented by the discharge valves of the intake and the inlet 
valves of the high-pressure cylinder, together with the friction 
in air passages and intercooler. This unavoidable loss should be 
minimized by making the valves, ports, and connecting passages 
of ample size. 

Second. As with single-cylinder compressors, the compres- 
sion line of each cylinder of most stage compressors departs but 

* This combined indicator card, which does not show all the minor irregularities 
in the lines, is from a Rand cross-compound compressor. It accompanies an 
article by F. A. Halsey, on " The Analysis of Air Compressor Indicator Diagrams," 
American Machinist, March 3, 1898, p. 158, and is reproduced here by permission. 



98 COMPRESSED AIR PLANT 

little from the adiabatic curve. Aside from the thermodynamic 
advantage of dividing the total compression between two or 
more cylinders, and thereby lowering the average and final 
temperatures, it is the intercooler that must be rehed on for 
the chief element in economical working. By its abstraction of 
heat the volume of air entering the second cylinder is reduced, 
so that PV"" ^'^ = constant becomes approximately PV = C, on be- 
ginning the second stage. But the compression line again rises 
rapidly from this point and continues not far below the adiabatic. 
Indicator cards from dry compressors which do not show approxi- 
mately this relation between the lines are always open to sus- 
picion. A leaky piston, for example, will lower the compression 
curve and make it appear that better work is being done than 
is really the case. 

In constructing and reading a combined indicator card from 
a stage compressor, the adiabatic line applying to the com- 
pression in the second cyhnder should be represented in its 
proper place (see Fig. 51). The complete graphic relation be- 
tween the several heat curves is thus set forth. 

Third. It is an advantage of stage compression that there is 
practically but one clearance space — that in the intake cylinder, 
and, as the air in this cylinder is at a low pressure, the resulting 
reduction in volumetric capacity is moderate, for the loss due 
to clearance is proportionately less for low than for high pres- 
sures. The piston clearance of the high-pressure cylinder 
cannot affect the volume of air delivered, because all the air 
discharged from the intake cylinder goes to the high-pressure 
cylinder and, barring leakage, must pass through it. 

The heating of the cylinder walls and pistons reduces some- 
what the working volumetric capacity of an air compressor, 
because, as the entering air is warmed, a smaller weight of it is 
taken into the cylinder at each stroke. Although the degree 
of this heating cannot be formulated, it is obviously less in a 
two-stage than in a single-cylinder compressor; for, aside from 
the effect of the intercooler, the smaller quantity of heat 
generated in each cyhnder is more efficiently dealt with by their 
respective water-jackets. 



CHAPTER VII 
AIR INLET VALVES * 

The design and working of the inlet or suction valves greatly 
influence the compressor's efficiency. That there are still 
wide differe!nces of opinion as to the best design is evidenced 
by the variety of types in use, and the lack of clearly defined 
distinctions as to their applicabihty under given conditions. 
In the older wet compressors, as the Dubois-Francois, various 
patterns of clack-valve were employed. Some are still used in 
Europe, like the elaborate, cam-controlled clack-valves of a 
compressor built by Schneider & Co., Creusot, France. For 
years poppet valves of numerous types held the field in the 
United States. They are furnished with springs, and are now 
actuated solely by difference of air pressure; though there have 
been examples of mechanically controlled poppets, as in the 
old Rand mechanical valve-gear. 

While poppets are favored for some kinds of service, several 
other forms of inlet valve have been introduced. Modifications 
of the Corliss steam valve, first used in the Norwalk compressor, 
have been adopted for compressors of many other makes, as the 
Sullivan, Nordberg, Ingersoll-Rand, Laidlaw-Dunn- Gordon, and 
Allis-Chalmers. Besides these are the so-called " thin-plate 
valves," discussed later in this chapter. 

Chief Requisite's of Inlet Valves: i. They must have a 
sufficient area of opening to permit free entrance of the air. 
2. They must open readily near the beginning of the stroke, 
with minimum resistance, remain open until the end of the stroke, 
and then close promptly. 

* This chapter is devoted chiefly to poppet valves and others which operate 
by difference of air pressure. For discussion of mechanically controlled inlet 
valves, see Chap. IX. 

99 



100 COMPRESSED AIR PLANT 

The point of stroke at which the inlet valves open depends 
on the piston clearance and terminal air pressure. Valves oper- 
ated mechanically are sometimes incorrectly designed or set, 
to open exactly at the beginning of the stroke or a fraction 
later; in which case the clearance air is first exhausted through 
the valves and then, as the piston advances, the outside air 
begins to enter. This being so, little or no clearance would be 
shown on the indicator card. Premature closing reduces the 
volume of intake air, and hence the volumetric capacity of the 
compressor. Its effect on the indicator card is to lower the 
compression line near the beginning of the stroke, so as to 
approach the isothermal curve and make it appear that the 
compressor is doing abnormally good work. 

As pointed out in Chap. Ill, although piston clearance 
reduces the volumetric capacity of the cylinder, it does not 
cause a corresponding loss of work, and it is of some benefit 
in assisting to overcome the inertia of the reciprocating parts 
of the compressor. Part of the work expended in compressing 
the clearance air is thus recovered, whereas, when the clearance 
air is exhausted by a premature opening of the inlet valves, 
the work represented by it is lost. The proper adjustment of 
spring-controlled poppets is a question of the strength of the 
spring, and since the effect of clearance varies with the terminal 
air pressure, the valves must be regulated accordingly. Any 
exhaust through the inlet valves is readily detected by the noise. 
When they are properly set, the compressor works more smoothly 
and the power consumed is sHghtly reduced. On the other hand, 
if the valves open too late in the stroke — due, for example, to a 
temporary reduction in working pressure — a little more power 
is required, this condition being shown by the slight drop in 
the re-expansion line at the point b (Figs. 45 and 51). 

Inlet Area. The total area of the inlet ports varies greatly 
in compressors of different makers. It is sometimes as small 
as 4 or 5% of the piston area, running to a probable maximum 
of 15%. As the proper area is a function of the piston speed, 
it may be made less for slow- than for high-speed compressors. 
However, in one of the Leyner 2-stage compressors, with a 22-in. 



AIR INLET VALVES 101 

intake cylinder and running at the moderate piston speed of 
390 ft., the intake port area is 14.2% of the piston area. To 
prevent excessive frictional resistance during the inflow of air, 
the inlet area, under average conditions and for ordinary forms 
of valve, should be not less than 8 or 10% of the piston area. 
But if poppet valves are unnecessarily large, their inertia becomes 
great; and if too numerous, there are more parts to care for, 
and valuable water-jacket area on the cylinder heads is sacrificed. 
In the two-stage, straight-line, '' Hurricane-inlet " compressors, 
of the Ingersoll-Rand Co., type AA-2, for cylinders from 15-in. to 
24-in. diameter, the inlet area of the intake cylinder averages 
13.2% of the piston area. For the high-pressure cylinders of 
the same compressors, the poppet valves have an average inlet 
area of about 11%. The inlet area of the duplex, two-stage 
compressors of the same builders, type " O-2," averages 13.6% 
of the piston area for both low- and high-pressure cylinders. 
In one type of the two-stage compressors of the Laidlaw-Dunn- 
Gordon Co. the percentage is 12-14. 

Poppet Inlet Valves. A common form is the mushroom 
valve, Figs. 52 and 53. While the total inlet area should be 
ample, there are two special requirements in the case of ordinary 
poppet valves: (i) the area of each valve must be moderate, 
or the valve will be too heavy, causing unnecessary injury to 
the seat, and by its inertia too much resistance to the control 
of the spring; (2) the lift must be small, to insure prompt opening 
and closure, and to reduce " chattering," as well as wear. For 
these reasons the total area required is furnished by a number of 
independent valves, generally from 4 to 8. 

The valve is of steel or bronze, with an easily removable 
bronze seat, the contact surfaces being ground true and the seat- 
ing preferably coned. The stem works in guides, forming part 
of the valve casing, which is screwed into the cylinder head 
so as to be readily removed for adjustment or regrinding. 
Brass springs are used to avoid the effects of corrosion, and 
must be easily compressible to allow the valve to open under a 
small difference of pressure. The springs must be of the best 
material and calibrated to present no more than the minimum 



102 



COMPRESSED AIR PLANT 




Fig. 52. — Norwalk Poppet Inlet Valve. ^ 




Ftg. 53. — Laidlaw-Dunn-Gordon Poppet Inlet 



AIR INLET VALVES 103 

requisite resistance, which varies from, say, 3 oz. to 8 or 10 oz. 
per sq. in. of valve area. 

Ordinary poppets are opened by atmospheric pressure from 
without, when a certain degree of rarefaction of the air inside 
the cylinder has been produced by the movement of the piston; 
in other words, when the difference of pressure, after the clear- 
ance air has re-expanded, becomes sufficient to overcome the 
spring resistance. The loss of volumetric capacity due to spring 
resistance, in terms of free air, is rarely less than 2-3%, and is 
often more. At sea-level a spring pressure of 5 oz. per sq. in. of 
valve area causes a loss of about 2%. The diagram. Fig. 54, 
shows the effect of spring resistance in reducing the volumetric 
capacity of a compressor at different altitudes, from sea-level 
to 15,000 ft. elevation. 

Spring-controlled poppets cause more or less irregularity 
in the entrance of the air, because, while the pressure of the 
outside air tries to open the valve, the spring tends to keep 
it closed. This often produces '' chattering " or '^ dancing " of 
the valves, and has led to the introduction of various mechanical 
devices for definitely controlling them, as noted later. As the 
springs lose their original elasticity, and undergo alterations 
in strength, they require regulation from time to time; adjust- 
ing nuts on the valve stems are generally provided. If the 
springs be too slack, chattering^ increases; if too tight, the 
valves will open late in the stroke, and the air filUng the cylinder 
may have a density less than that of the atmosphere. Aside 
from the spring resistance, the rate of inflow of the intake air 
varies with the speed of the piston. When its speed is greatest, 
at the middle of the stroke, the rate of inflow is at the maximum. 
While it is moving slowly, near the beginning and end of each 
stroke as the crank turns its centers, the relatively small negative 
pressure is insufficient to open the valves and keep them open 
against the springs. 

The total valve resistance, including that due to throttling 
of the intake air and friction in passing through the ports, is 
kept as small as practicable, but can never be entirely eliminated. 
With some forms of inlet valves, other than spring poppets, the 



104 



CO:\IPRESSED AIR PLANT 



Altitude Above Sea Level, Feet 



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Fig. 54. — Effect of Valve Spring Resistance on Volumetric Capacity of Compres- 
sors. {Eng. News.) 



AIR INLET VALVES . 105 

resistance may be very small. Its usual effect is shown by the 
diagram, Fig. 55. There is generally sufficient resistance to 
keep the admission line, AC, at an appreciable distance below 
the atmospheric line, DE, throughout the stroke; the amount 
of loss from this cause is measured by the area of the indicator 
diagram lying below the atmospheric line. If the inlet area 
be too small or the valves poorly designed, the negative pressure 
may amount to i or 2 lbs. per sq. in. The point B, where the 
com.pression line crosses the atmospheric line, is the point of the 
stroke which must be reached by the piston before any useful 
work is done, and the volume passed through in travelling from 



_pelivTres6.=65Lbi / 



Speed=90 Revs. 

"V0l.Effic.-=88^ 



Fig. 55. 

A to B represents the loss in volumetric capacity from this cause. 
The total loss of volumetric capacity, including that due to 
piston clearance, is represented by the length of AB + CE, 
and the volumetric efficiency of the compressor is equal to the 
length of the line BC, divided by the total length of the diagram. 
Notwithstanding certain inherent disadvantages, the poppet 
valve is widely used, for both inlet and discharge. It is simple, 
easily regulated, and in case of leakage, due to cutting or unequal 
wear of the seating surfaces, is readily removed and re-ground. 
In stage compressors it is often used for the high-pressure 
cylinders, even when some other type is preferred for the low- 



106 



COMPRESSED AIR PLANT 



pressure. Poppets must be kept clean; they not infrequently 
cause trouble by sticking in their seats due to the accumulation 
of gummy oil, or they may be clogged by deposit of carbonaceous 
matter from decomposition of the lubricant, produced by 
excessive heating of the cylinder. 

One form of Norwalk two-stage compressor has a special 
poppet inlet valve, for use when it is desired to employ air 
at two different pressures, obtained from a single compressor. 
In stage compression, though the air is actually produced at 
two pressures, of say 25-30 and 80-100 lbs. respectively in 
the low- and high-pressure cylinders, yet, if a part of the volume 




Fig. 56 — " Skip- Valve." Norwalk Iron Works Co. 



delivered by the intake cylinder be drawn from the intercooler, 
the high-pressure cylinder cannot work satisfactorily. The 
air remaining in the intercooler expands to a lower pressure 
before going to the high -pressure cylinder, so that the ratio 
of compression in this cylinder is increased, and the heat gener- 
ated in it rises to a correspondingly higher degree. The rise in 
temperature produced by a considerable increase in the ratio of 
compression would prevent proper lubrication, and the con- 
ditions might be favorable for an explosion in the cylinder 
(Chap. XIV). This difhculty is met by using '' skip- valves " 
(Fig. 56) as inlet valves of the high-pressure cylinder. They 



AIR INLET VALVES 107 

open, and remain open, whenever the high-pressure inlet air 
falls below the normal, by reason of having drawn off a portion 
of the air from the intercooler. The high-pressure cylinder is 
thus partly unloaded, since the air entering at each stroke is 
returned to the intercooler. The skip-valve is a mushroom 
spring-poppet DE, carried in the guides AA. Above the valve 
is a small spring- controlled plunger B, the space below which is 
occupied by air at intercooler pressure. When this pressure 
falls below that for which spring C is set, the plunger advances 
and forces open the inlet valve, holding it open until the inter- 
cooler pressure rises sufficiently to cause the plunger to recede. 
The valve is then free to work in the usual manner. The action 
of the valve thus adjusts itself to the varying pressure of the 
intake air coming from the intercooler. 

Corliss Air Valves. As these are mechanically controlled, 
details are given in Chap. IX. 

Ingersoll-Rand "Hurricane-Inlet" Valve, a modification of 
the old " Piston-Inlet " valve, is shown by Figs. 57 and 58. 
It has in turn been largely displaced by " plate valves," noted 
below. The piston is hollow and has a hollow back piston-rod 
for admitting the air. There are two large, ring-shaped valves 
(one in each face of the piston), of T cross-section. They are 
held in place, without springs or other connection, by guide- 
plates bolted to the piston faces. Their play is limited by these 
guides (see Fig. 58), which contain a series of circular ports, 
furnishing additional area for the passage of air. The valves are 
readily taken out for regrinding by removing the guide plates. 
At the beginning of each stroke, as the piston reverses, the valves 
alternately open and close by their own inertia. The valve 
in that face of the piston which is toward the direction of move- 
ment is always closed, while the other is open for the passage 
of the air entering through the hollow rod into the cylinder 
behind the piston. Because of the large diameter of the valves 
their throw, or Hft, is small — in ordinary compressors say f in. 
The area of the inlet tube is usually 13-14% of the piston area. 
Though the actual port area of the valve is less than this (about 
8%), the velocity of flow is moderate, because, instead of having 



108 



COMPRESSED AIR PLANT 




AIR INLET VALVES . 109 

a group of inlet valves, the inlet is concentrated in a single 
annular opening. 

The space in each cylinder head that would otherwise be 
occupied by inlet valves is utilized for additional water-jacket 
area. It has been objected that, since the inlet tube and pistpn 
are necessarily heated, the temperature of the intake air is 
raised in its passage into the cylinder; and that therefore the 
weight of air in the cyHnder is relatively less than if it had 
entered by a more direct path. But, as the air enters in a 




Fig. 58. — IngersoU-Ra.id " Hurrlcane-lulet," Enlarged Section. 

single large stream, instead of being divided into comparatively 
small areas of flow, it can absorb little heat until it reaches 
the valve, because only a thin film of the rapidly moving air in the 
inlet tube is in contact with the hot metal. Some heat is 
absorbed by the air in passing in the thin sheet through the 
valve port in the piston face; but thermometric observations, 
taken inside the inlet tube and piston, at speeds of 40 and 120 
revs, per min., indicate a rise of not over 5° F. at the lower 
speed and even less at the higher. It is unlikely that better 
results are obtainable from either poppet or Corliss valves. 

Plate Valves. For many years, " thin-plate " valves of 
several forms have been occasionally used; for example, the 



no COMPRESSED AIR PLANT 

Leyner (U. S.), the Guttermuth, and the Riedler (Europe), 
all now nearly obsolete. Since about 1913, interest in 
this t>^e of valve has revived, and it is now used by 
a number of compressor builders (IngersoU-Rand, Chicago 
Pneumatic Tool Co., Laidlaw-Dunn-Gordon, Sullivan, AlUs- 
Chalmers, Robey, and Walker), for inlet or discharge valves, 
or both. 

Plate valves appear to be satisfactory, though some think 
they have not yet been thoroughly tested as to endurance at 
the higher compressor speeds permitted by their small inertia. 
They probably lower the cost of manufacture. 

The Guttermuth valve is a rectangular plate of thin steel 
with a grid seat. One side of the plate is coiled in a spiral, 
around a stationary spindle, the inner edge of the spiral being 
inserted in a longitudinal groove in this spindle. By placing 
several valves side by side any desired area of opening can be 
furnished. To avoid the harmful efi'ects of inertia, the valves 
are very thin, with sensitive springs, and by so arranging them 
that the current of air in passing through the valve into the 
cylinder undergoes but slight changes of direction, serious 
eddying of the air around the edges of the plate is pre- 
vented. 

Leyner Annular Valve is shown by Figs. 59 and 60. Though 
no longer made, some of the compressors using it are still in 
service. Fig. 59 is a longitudinal section through the adjacent 
ends of the low- and high-pressure cyUnders of a straight-line, 
two-stage compressor, indicating incidentally the circulation 
of the air through the intercooling tubes. The inlet and dis- 
charge valves differ only in size. The valve (Fig. 60) is a thin 
steel plate cut in a peculiar form. The outer, or seating portion, 
is a narrow annulus, with two arc-shaped strips terminating 
in a central ring, which is locked to the cylinder head by a nut 
encircHng the piston-rod. The arc strips, connecting the seating 
part of the valve mth the fixed center serve as springs. There 
is but one inlet and one delivery valve at each end of the cylinder. 
The inlet ports, DD (Fig. 59), 4 in number, are curved, slot-like 
openings. There are 6 similar but smaller discharge ports, 



AIR INLET VALVES 



111 




> 



Q 

a 
a 

la 






bo 

a 

O 

C/3 



C/2 






t 

O 









o 



112 



COMPRESSED AIR PLANT 




Fig. 6o. — Leyner Annular Inlet Valve. 




Fig. 6i. — Laidlaw-Dunn-Gordon Air Cylinder, with Valves in Place 



AIR INLET VALVES 



113 



EE. Total area of inlet ports is about 14%, and of discharge 
ports, nearly 9% of the piston area.* 

Laidlaw-Dunn-Gordon " Feather Valve." Fig. 61 shows 
the cylinder, with the inlet and discharge valves, which are 





Fig. 62. — Laidlaw-Dunn-Gordon " Feather " Valve, with Cover Removed. 





.^ _^-, ,^v#sisjas^i ' 



Fig. 63. — Laidlaw-Dunn-Gordon " Feather " Valve, Assembled. 

alike and interchangeable; Fig. 62, the valve with cover 
removed; Fig. 63, the bottom and top views of the assembled 
valve. Figs, i and 24 (Chap. II) include sectional views of 

* Tests have shown an intake pressure loss of o.g oz. On a card with a 20- 
scale spring, this would be represented by a difference of only 0.003 ^^- between 
the intake and atmospheric lines. 



114 



COMPRESSED AIR PLANT 



the Talve, with its setting in the cyHnder. The valve consists 
of 5 thin rectangular steel strips, the middle portions of which 
are free to Hft from the seat to permit the passage of air. When 
the valve opens, the ends of each strip remain in contact with 
the seat, being retained in place between the valve cover and 

seat, though not held rigidly. 
As the steel strips are very 
Hght. the destructive effects 
of inertia are minimized. 
Fig. 24 shows the yoke and 
set-screw by which the valve 
cover is held in place. 

Ingersoll-Rand Plate Valve 
(Figs. 64, 65, 66). Referring 
to Figs. 65, 66, the valve F is 
a perforated disk of thin 
plate. The outer or seating 
portion is connected to the 
central ring X (firmly at- 
tached to seat A by stud 
bolt B) by the arms MM. 
Thus the valve can not turn 
on its center, but always 
rests in the same position. 
Arms M are ground thinner 
than the valve itself, to se- 
cure elasticity. Thickness 
of a medium size valve, 0.07 
in.; diameter varies from 
4.5 to 15 ins. The four light springs H of the cushion plate or 
buffer hold the valve on its seat A, when at rest, against the 
slight resistance of MM. When the required air pressure is 
reached, the valve rises against springs H to its full opening. 
Fig. 66 is a section of the valve assembled. The valve and 
cushion spring are separated by the washers G and E, and 
are attached to the seat by the bolt B. As there are no rubbing 
guides, lubrication is unnecessary. The port area is large; 




Fig. 64. — Cylinder and Intercooler of 
Ingersoll-Rand Class " ORG " Com- 
pressor, with Plate Valves. 



AIR INLET VALVES 



115 



height of lift is from 0.08 to 0.14 in., according to size of valve. 
Weight of the valve is about one-third that of an equivalent 
poppet valve. The inlet and discharge valves are identical. 




CUSHION PLATE 



Fig. 65. — Parts of the Ingersoll-Rand Plate Valve. 

This valve is now being extensively used in the compressors of 
the Ingersoll-Rand Co. It somewhat resembles the older 
Leyner annular valve (Figs. 59 and 60). 




Fig. 66. — Section of Ingersoll-Rand Plate Valve, Assembled. 



Sullivan Plate Valve {'^ end rolling finger- valve ") was applied 
in 191 7 to the angle- compound compressors (see Fig. 25, Chap. 
11) of these makers. Fig. 67 shows the design of valves and 



116 



COMPRESSED AIR PLANT 



seats in the low-pressure cylinder, for both inlet and discharge; 
Fig. 68, the position of the valves in the end of the cyHnder. 
The valve consists of a group of thin, flat strips {" fingers "), 
of spring steel, mth grid seats. Over the fingers is a correspond- 
ing row of curved guard plates. As the fingers open they roll 
against their guard plates from the stationary end towards the 
tip, and conversely on closing. 

In the high-pressure cylinder, the inlet and discharge valves 
are also of the plate type, but are placed in cages set radially 




DISCHARGE VALVE AND 
BACKING OR GUABD PLATE 



NLET 

GUARD 
PLATES 



FINGERS OF INLET VALVE 

Fig. 67. — Sullivan Plate Valve, Low-Pressure Cylinder, Angle-Compound Com- 
pressor, Class " WJ-3." 



around the periphery of the cylinder heads. There are two 
valves in each cage, each consisting of two fingers running 
lengthwise of the cage. 

Arrangements for Conveying Inlet Air to the Compressor. 
The colder the intake air the smaller the volume occupied by a 
given weight of air taken into the cylinder, and the greater the 
volumetric output. ^Ir. Frank Richards states:* " The volume 
of air at common temperatures varies directly as the absolute 

*" Compressed Air," p. 55. 



AIR INLET VALVES 



117 



temperature. With the air supply at 60° its absolute tem- 
perature is 521°, and its volume will increase or decrease 



521 



for each degree of rise or fall of temperature. Therefore, if in 
securing the supply of air we can get a difference in our favor 
of 5° . . ., we accomplish a saving of about 1%. If a difference 
of temperature of 10° can be secured 2% is saved," practically 
without cost. The practice of taking air from the engine-room 
is a common one at mines, and is bad not only because such air 
is usually heated to a considerable degree, but is apt also to be 
charged with dust, which causes unnecessary wear of valves 





INSIDE OF END WALL 
OF CYLINDER 



OUTSIDE OF END WALL, 
OF CYLINDER 



Fig. 68. — Sullivan Plate Valves in Place on Inner Head of Low-Pressure Cylinder, 
Angle- Compound Compressor, Class " WJ-3." 



and piston. Fresh air should be taken, preferably from some 
point outside of the building. A box or pipe of wood is better 
than one of iron, because of the smaller conductivity of wood. 
Its cross-section should be sufficient, say, at least one-half the 
area of the cylinder, to avoid loss from friction. To make such 
a connection conveniently, the inlet valves should be enclosed in 
an external air chest on each end of the cyHnder. Compressors 
having a single inlet valve are well adapted for making this 
arrangement. Care should be taken to prevent the entrance 
of dust, leaves, or rubbish. If the inlet is open, particles floating 
in the air may be drawn in and obstruct the valves or injure 



118 COMPRESSED AIR PLANT 

their seats and the working surfaces of piston and cyhnder. 
By building a conduit from the outside of the compressor house 
to the air box enclosing the inlet valves, a greater saving can.be 
effected in winter than in summer, but even in warm weather 
some advantage is gained, especially if the conduit opens on the 
north side of the building, out of reach of the sun's direct rays^ 
and is carried vertically to some height above the ground. 
In the Ingersoll-Rand " Hurricane inlet," the outer end of the 
hollow back piston rod is inclosed by an intake duct (Fig. 57). 



CHAPTER VIII 
DISCHARGE OR DELIVERY VALVES 

The conditions affecting the action of the discharge valves of a 
compressor are wholly different from those governing the inlet 
valves. While the latter must open under very small differences 
of pressure, the discharge valves are subjected to a heavy pres- 
sure on both sides. Furthermore, owing to irregularities in 
the use of the air, the receiver pressure usually fluctuates consider- 
ably, so that the point of the stroke at which the discharge 
valves open cannot depend solely on the conditions, as to the 
ratio of compression, etc., under which the compressor is work- 
ing. The time of opening must depend also on the relation 
between the variable pressures in cylinder and receiver. Hence„ 
the sensitiveness of operation essential in inlet valves is unneces- 
sary for discharge valves. The chief requirements are that they 
shall be strong, free to open when the cyHnder pressure exceeds 
that of the receiver, have strong springs so as to close promptly 
at the end of the stroke, fit accurately on their seats, have large 
free air passage around the rim of the valve, and ample guiding 
surface to insure accurate alinement in their movements. Delay 
in closing, or leakage between valve and seat, are far more 
serious than for inlet valves, because these defects are equivalent 
to an increase of piston clearance and consequent reduction 
in the volumetric capacity of the cylinder. The leakage of 
even a small quantity of compressed air back into the cylinder 
is equivalent to the loss caused by an abnormally large clearance 
space. Compared with inlet valves, therefore, discharge valves 
afford a relatively limited field for innovation or improvement. 

Poppet Discharge Valves. Except the " plate valves " 
(Chap. VII), and a few designs in wh^ch mechanical control is 
introduced (Chap. IX), nearly all discharge valves are cup 

119 



120 



COMPRESSED AIR PLANT 



shaped poppets, with internal springs. Though varying in 
details, they are represented fairly by Figs. 70 and 71. Other 
designs are shown in the sections of air cylinders illustrated in 




Fig. 69. — Ingersoll-Rand " Imperial " Poppet Discharge Valve. 

Chaps. II and VII. They are more rarely of the mushroom type, 
like that in Fig. 69, because these may not afford sufficient 
guiding surface. Valves are of steel or bronze, with bronze 




Fig 70. — Laidlaw-Dunn-Gordon Poppet Discharge Valve. 

seat. To assist keeping them tight, the seating surfaces are 
coned. A group of poppets are commonly employed, to avoid 
making them of large size and weight. Under their high work- 



DISCHARGE OR DELIVERY VALVES 



121 



ing pressure, the inertia of heavy valves causes destructive wear. 
Each valve should be readily accessible for adjustment, regrind- 
ing, or renewal. They are therefore covered by caps screwed 
into the cylinder head; or, in some makes, by cover plates on 
the valve chamber. 

Cataract-Controlled Poppets. In these the valve not only 
has a spring, but its action is further modified by attaching 
the valve stem to the piston of a small cataract cylinder, 
containing air or oil, to ease its movement and avoid hurtful 
shocks.* These valves are employed to a considerable extent 




Fig. 71. — Norwalk Poppet Discharge Valve. 

in Europe, but not in the United States. Some are satis- 
factory for slow piston speeds; for high-speed compressors 
they do not work with sufficient promptness to prevent '' slip " 
or leakage of compressed air back into the cylinder. Their 
chief object may be attained in another way, as by the partial 
control of an accompanying Corliss valve (see Fig. 75, Chap. IX). 
Riedler Discharge Valve (Fig. 72) represents one of several 
patterns employed in the Riedler compressors. It is a light, 
cylindrical valve A, with packing rings D. The cylinder in 
this case is vertical, and the piston L carries at its periphery 
plate P, held in place by stud N and spring M. When closed 
the valve seats on plate E, being held against it by the air 

* Cataract-controlled poppets are employed by some European compressor- 
builders for both inlet and discharge. 



122 



Compressed air plant 



pressure in the discharge passage acting on the under side of the 
upper flared end. In this position the round air ports near the 
lower edge of the valve are closed by the valve guide, at CC. 
As the piston advances, and when the cylinder pressure exceeds 
that in the receiver, the valve is opened by the air pressure on 
the upper side of the flared end. This movement of the valve is 
cushioned by the air trapped above the guide BB. On reach- 
ing the end of the stroke, the plate P, on the piston, strikes 
the lower edge of the valve and closes it against its seat E,. 
the shock being cushioned by springs F and M. The double 




Fig. 72. — "Express" Poppet Valve, Riedler Compressor. 

cushioning, in both opening and closing, tends to durability; and 
moreover, when plate P strikes the valve, the crank is nearly on 
its center, so that the piston is moving very slowly. Mechan- 
ically controlled air-valves of the Riedler design are described 
in Chap. IX. 

** Thin-Plate '' Discharge Valves are being used in some 
recent types of compressor. They are identical with the corre- 
sponding inlet valves. For examples, see Figs. 61-68, Chap. VII. 

Several other forms of discharge valve are noted in Chap. IX, 
in connection with mechanically controlled valve motions. 

Discharge Area for Air Cylinders. The volume of air to be 
discharged having been reduced by compression to a small 



DISCHARGE OR DELIVERY VALVES 123 

fraction of the volume at atmospheric pressure, it might appear 
that the total discharge area could be made much smaller than 
the inlet area, without causing excessive frictional resistance. 
But the compressed air must be forced out of the cylinder 
in a relatively short period of time. While the inlet valves are 
open throughout nearly the entire stroke, the delivery must 
take place while the piston is making, say, the last third of 
the stroke. Therefore, in a compressor of ordinary design, with 
several poppet inlet and discharge valves, the total discharge 
area should be about equal to the inlet area, provided the 
piston speed be moderate. When the inlet area is concentrated 
in a single valve (for example, the Inger soil- Rand '' Hurri- 
cane-Inlet ") , the discharge area is about double the inlet 
area, though this relation varies in cylinders of different sizes, 
being proportionately greater in the larger compressors. Other 
things being equal, the discharge area should increase with the 
piston speed. For a speed of 350 ft. per min., the discharge 
area may be, say, 10% of the cylinder area; for speeds of 450-600 
ft. per min., 15%. In some compressors, the discharge area 
is as small as 8.5-9.5%. 

The above considerations apply also to the passages through 
which the air passes from the discharge valves to the pipe 
leading to the receiver. In some designs these are too restricted 
to permit a free flow of the air. The velocity of discharge should 
be kept low to minimize frictional resistance; otherwise, during 
the delivery period the pressure in the cylinder will rise momen- 
tarily above the normal, and then drop back after the air has 
passed out to the receiver. This causes loss of power and 
unnecessary strains on the moving parts of the compressor. 
The amount of this loss is represented by the irregular area 
of the air card which lies above a horizontal hne drawn through 
the point corresponding to the pressure at the end of delivery 
(Fig. 55). When the discharge valves first open, the piston 
is moving at a high velocity, and equilibrium with the receiver 
pressure is only attained as this velocity decreases toward the 
end of the stroke^ 



CHAPTER IX 
MECHANICALLY CONTROLLED \'.\LVES AND VALVE MOTIONS 

Mechanical Control for Inlet Valves. The disadvantageous 
features of inlet valves that depend primarily for opening and 
closing upon difference of air pressure have led to the intro- 
duction of numerous mechanically controlled valves. By their 
use fewer valves are required, because they may be made larger 
and have a higher lift. They are controlled by being in some 
way connected with the rotary or reciprocating parts of the 
compressor. By thus modifying the effect of the valve spring, a 
prompter opening is secured, so that the compressor takes more 
nearly a full cylinder of air at each stroke. 

In some designs the connection between the valves and their 
operating mechanism is positive and fixed for any one setting 
of the valves, which are timed with respect to the piston stroke, 
to open at the instant the clearance air has been re-expanded to 
atmospheric pressure, and to close at the end of the stroke. In 
other designs springs are used in connection with the controlHng 
mechanism, thus allowing for variations in working conditions, 
as well as for inaccuracies of adjustment or slight derangements 
caused by wear of parts. Still other valve motions exert a 
partial control, which, within narrow limits, leaves the valve 
free to act under difference of air pressure inside and outside of 
the cylinder. 

As a rule, the inlet valves only are positively controlled, 
and in most cases the Corliss type of valve is used. While 
mechanically controlled valves may be employed for the intake 
cyUnders of stage compressors, they are not suitable for the 
high-pressure cylinders; the inlet valves of these are subjected 
to heavy pressure on both sides, and are best allowed to opep 

124 



MECHANICALLY CONTROLLED VALVES 125 

and close solely under the difference between these pressures, 
which is more than sufficient to produce prompt action of the 
valve at the proper time. Poppet valves are therefore generally 
used for this service. 

Mechanical Control for Discharge Valves is generally un- 
satisfactory, because of the fluctuations of receiver pressure. In 
attempting to open them by a positive mechanical movement, 
at a fixed point of the stroke, two cases may occur: (i) in event 
of a drop in receiver pressure below the normal, the valves 
and their controlling mechanism would be subjected to a heavy 
strain, before the point of opening is reached, due to excess of 
cylinder pressure; and, (2) if the receiver pressure should rise 
above normal, the valves would be held forcibly on their seats, 
by the excess of receiver pressure, after being released by the 
controlling gear. In either case, derangement or breakage of 
some part of the controlling mechanism may occur. 

To allow the discharge valves to adjust themselves auto- 
matically to the varying conditions they must have some degree 
of freedom as to their time of opening. Only a partial control 
is practicable for any service in which the consumption of air is 
variable. Moreover, Corliss valves as used for compressors 
do not serve well for discharge valves where the ratio of compres- 
sion is greater than, say, three or three and one-half; because they 
must then be set to open too late in the stroke to permit a free 
discharge. This applies to ordinary single-stage compressors, 
as well as to the high-pressure cylinders of two-stage machines. 
A number of devices have been introduced for dealing with these 
conditions; for example, relief valves working in conjunction 
with mechanically operated discharge valves. 

Valve Motion of Norwalk Compressor. An adaptation of the 
Corliss valve gear is used for the low-pressure cylinder of this 
compressor (Fig. 73). One inlet and one discharge valve are 
set in chests at each end of the cylinder. Poppet valves are 
employed for high-pressure cylinder (as shown in Fig. 6, together 
with cross-sections of the low-pressure Corliss valves). In Fig. 
73, the main valve-rod a is driven by a drag- or return-crank b, 
mounted on the fly-wheel crank-pin. The rod is pin-connected 



126 



COMPRESSED MR PL.\XT 




MECHANICALLY CONTROLLED VALVES 127 

to a short lever c, on the spindle of the forward inlet valve, and 
from this lever a link d passes to a corresponding connection 
with the inlet valve at the other end of the cylinder, the parts 
being so adjusted that one valve opens as the other closes. A 
positive movement of the valves is thus obtained. Cams, / / 
and g g, for operating the discharge valves, are mounted on the 
respective inlet and discharge valve spindles, and form part of 
the short levers c. As each inlet valve oscillates, its cam rolls 
upon that of the discharge valve above it, each pair of cams being 
shaped so that the discharge valve is opened full when the 
cylinder pressure equals that in the discharge passage outside. 
Then, at the end of the stroke, when the cams move in the 
opposite direction, still rolling upon each other, the discharge 
valve is closed without shock by the connecting link e. This 
link is made of two telescoping parts, on the principle of a dash- 
pot, thus allowing freedom of movement for dealing with variable 
receiver pressure. 

A number of other compressor builders have adopted modi- 
fications of the Corliss valve gear for the air cylinders. 

Nordberg Valve Motion. For single-stage compressors the 
inlet valves are of the Corliss type, poppets being used for 
discharge (see Fig. 42). The inlet valves are operated positively 
from a triple wrist-arm, on the side of the air cylinder, driven 
by an eccentric on the crank-shaft. Connecting links pass 
from the wrist-arm to the valve levers. The lap of the valves 
can be altered by slightly shifting the angular position of the 
lever with respect to the valve spindle on which it is mounted. 
This is done by adjusting screws on the hub of the valve lever, 
the correctness of the valve motion being unaffected. 

The stage compressors have similar inlet valves, a modified 
Corhss valve, with double ports, being used for discharge (Fig. 74). 
This valve is shown open on the right and closed on the left-hand 
end of cylinder. An opening in the center of the valve allows 
the air to discharge on both sides. In the axis of each Corhss 
valve are set a series of spring poppets, which act as relief valves 
when the receiver pressure falls below the normal. As shown by 
the cut the water- jacketed area on the cyHnders is unusually 



128 



COMPRESSED .\IR PLAXT 



large, jackets being applied to the heads and around the valves, 

as well as on the cylinder barrels. 

•I 

Another form of Xordberg valve gear is used in compressors 
intended for constant speed, either steam- or belt-driven. (Figs. 
107, 108.) \\Tiile the general, construction of the air cyhnders 
is the same as shown in Fig. 42, the inlet valves have a releasing 
mechanism. \Mien the air pressure is normal the valves are 
operated as usual by T\Tist-plate Hnks. But when the pressure 
increases the inlet valves are released and held open until the 
pressure drops; that is, the compressor is unloaded for the time 




Fig. 74. — Air Cylinder of Xordberg Stage Compressor. 



being, useful work ceasing. The release is effected by knock-off 
cams, similar to those used for Corliss steam valves, these cams 
being operated by a loaded plunger to which the compressed air 
is admitted when the pressure exceeds the normal. The com- 
pressor is thus self-regulating within small Hmits. For duplex 
compressors, added delicacy of regulation is obtained by design- 
ing the knock-off cams to unload in four successive steps, accord- 
ing to the variation in air pressure. 

Laidlaw-Dunn-Gordon Valve Motions. Several types of 
mechanical valve motion were formerly made by these builders. 
A number are still in operation, but the makers now employ the 



MECHANICALLY CONTROLLED VALVES 



129 



" feather valves," shown in Figs. 6i, 62 and 63, Chap. VII. 
In the '^ Cincinnati " valve gear (Fig. 75) a single CorHss valve, 
at each end of the cyhnder, serves as both inlet and discharge. 
The valve at the right-hand end of the cylinder is in position for 
admitting air, while that at the left is open for discharge, ,the 
corresponding inlet being closed. A large poppet is set vertically 
just above each Corliss valve. The latter is timed to open the 
port early enoup;h in the stroke to leave the poppet free to rise 
whenever the cylinder pressure exceeds receiver pressure. At 




Fig. 75. — " Cincinnati " Valve Gear, Laidlaw-Dunn-Gordon Compressor. 

the end of the stroke the Corliss valve takes its inlet position 
(right hand of cut), and at the same time, by shutting off the 
discharge, confines a little compressed air in the passage under the 
poppet. This air acts as a cushion, allowing the poppet to seat 
itself slowly and without shock, during the return stroke of the 
piston. Fixed mechanical control is thus exerted at three points: 
opening and closing of the inlet, and closing of the discharge. 

Allis-Chalmers Valve Motions. Fig. 76 shows Corliss inlet 
valves, operated from a triple wrist-arm driven by an eccentric 
on the fly-wheel shaft, 



130 



COMPRESSED AIR PLANT 



The discharge valves (five in number for ordinary sizes of 
compressor) are spring-poppets of the cup form. Another de- 
sign of discharge valve, formerly employed by these builders, 
consists of a cup poppet, without a spring, which is permitted 
to open freely, but is closed positively by a plunger actuated 
from a separate wrist-plate and eccentric. A single valve is 
placed in each cyhnder head. The plunger, carried by ex- 
terior guides, works within the valve and is so timed that it 
forces the valve to its seat just at the end of the stroke. On 




Fig. 76. — Standard Air Valve Motion. AUis- Chalmers Co. 



the return stroke the plunger recedes, while the valve is held 
on its scat by the receiver pressure until the cylinder pressure 
rises sufficiently to open it. In closing the valve, the plunger 
is cushioned on the air in the cup valve, which is thus seated 
without shock. 

Another design uses Corliss valves for both inlet and dis- 
charge. The time of closing the discharge is adjusted for the 
maximum working pressure. To allow for variations, small 
auxiliary spring poppets act as relief valves. These open 
freely when the receiver pressure falls below the pressure for 
which the Corliss valves arc set. 



MECHANICALLY CONTROLLED VALVES 



131 



In a later design (Figs. 14 and 43), 9 low-lift plate valves 
are used for discharge at each end of the cylinder. Light springs 
seat the valves promptly at the beginning of the intake stroke. 
No lubrication is required. 

Sullivan Valve Motions (see Figs. 5, 7 and 8, Chap. II). 
In the straight-line, compound, two-stage compressors, Cl^ss 
'' WC " (Figs. 7 and 8), the low-pressure cylinder has cup 
poppet inlets and Corliss discharge valves; the high-pressure 




Fig. 77. — Sullivan Air Cylinder, Corliss Inlet and Poppet Discharge Valves. 



cylinder has CorHss valves for inlet and poppets for discharge. 
The two-stage '' WB-2 " Class (Fig. 5), has a similar valve 
motion. The Corhss valves are operated from an eccentric 
pin (drag-crank) on the main crank-pin. Fig. 77 shows a 
design for a single-stage compressor. In an older pattern of 
these makers, the discharge valves are set radially around the 
upper part of the cyKnder-head castings. The piston clear- 
ance can thus be made very small, and the valves, placed in 
removable seats, are surrounded by the water-jackets. 



132 co:mpressed air plant 

Other Mechanically Controlled Valve Motions are employed 
in a number of compressors, as the Franklin, Clayton, Rix, 
American, etc. With one exception, all of the mechanically 
.controlled air valves referred to in this chapter are adaptations 
of the Corhss steam valve. A wholly different t}pe, however, 
is the 

Riedler Valve Motion. This has undergone radical modi- 
fications since its introduction, about 1888. Figs. 78, 79 and 80 
show the most recent design. Though no longer made in this 
country, some Riedler compressors are still in use here, and the 
valve motion is an interesting one. Mechanical control is 
exerted through a wrist-plate A, operated by an eccentric 
on the fly-wheel shaft. Back of A is a shding plate B, to which 
the links EE are pinned. Plate B is caused to leciprocate, 
through a distance equal to the lift of the valves, by two cams 
cut on the periphery of the wrist-plate and working against studs 
set in B. The motion thus transmitted through Hnks E 
oscillates the rock-shafts D, and throws the forked levers C 
(Fig. 79), which control the closure of the valves. 

The four valves, two suction and two dehvery, are similar 
in design. They are double-seated poppets, the air passing 
within the seating ring as well as around its periphery. The 
valve stem passes through a stufhng-box in the cylinder head 
into a bonnet bolted on outside. Within the bonnet is a dash- 
pot, the piston of which is attached to the valve stem. 

The inlet valve F, Fig. 79, operates as follows: At the begin- 
ning of the stroke the forked lever C is depressed by the rock- 
shaft and link, but leaves the valve free to open, its movement 
being steadied by the dash-pot piston G. The resistance 
presented by this piston is regulated by the screw H. For 
ordinary sizes of compressor the Hit of the valve is i in., giving 
a large area of opening. (The io}-in. valve shown in the cut, 
which is for the low-pressure cylinder of a 24-in. and 38-in. by 
48-in. compressor, has an area of 45 sq. ins.) Toward the end 
of the stroke the forked lever brings the valve gradually nearer 
its seat, as the piston velocity decreases. In completing its' 
movement at the end of the stroke, the lever forces the valve 



MECHANICALLY CONTROLLED VALVES 



133 







o 
o 



> 



t3 

f5 






134 



COMPRESSED AIR PLANT 




MECHANICALLY CONTROLLED VALVES 



135 




be 
S- 

> 

bO 
d 




*f5 



O 

OO 



O 



136 COMPRESSED AIR PLANT 

upon its seat promptly. The valve thus attains its maximum 
Uft in the middle of the stroke, when the velocity of the in- 
flowing air is gieatest, and is brought nearer its seat as the flow 
diminishes, so that closure is completed instantaneously at the 
proper time. 

The control of the delivery valve is similar, though the details 
of its bonnet, dash-pot, and forked lever are different (Fig. 80).* 
At the proper point of stroke the lever is depressed, so that the 
valve is free to open when the air pressure in front of the advanc- 
ing piston reaches receiver pressure. Then, as the velocity of 
outflow diminishes toward the end of the stroke, the valve 
approaches its seat, and closes promptly the instant the stroke 
reverses. 

Koster valve is of the piston type. It is employed by Bailey 
& Co., Manchester, England, and several compressor builders 
on the continent of Europe. The valves, both inlet and dis- 
charge, are large in area and mounted on a longitudinal spindle 
deriving its reciprocating motion from an eccentric on the 
crank-shaft. The opening and closure of the inlet valves are 
positive, but the delivery port is opened by an independent 
poppet, encircling the spindle and provided with a hght spring. 

* The deliver}' valve in the cut is the same size as the inlet valve in Fig. 79, 
but is designed for a smaller compressor, 15" and 24" X 3^"- 



CHAPTER X 
PERFORMANCE OF AIR COMPRESSORS * 

The duty of air compressors maybe designated in three ways: 
First. A useful standard of rating for ordinary purposes 
is the volumetric output, in terms of cubic feet of free air com- 
pressed per minute to a given pressure. The theoretical output 
is found by multiplying the net piston area in square feet by 
the distance travelled by the piston in feet per minute. The 
actual output is less than this on account of losses due to leaks, 
clearance, induction of warm air, friction of inlet valves, etc. 
In a well-designed compressor an allowance of 8 to 12% will 
cover these losses, which must not be confounded with the work 
required to overcome the friction of the compressor, and the 
added work due to the heating of the air while being compressed. 
The work losses are dealt with later. 

Having found the volumetric capacity of the compressor, 
the volume V of this air, at any given pressure P', is calculated 

VP 

by the formula : V' = ^^. where 

V = initial volume of free air, cu.ft. ; 
P = normal absolute pressure of atmosphere (14.7 lbs.); 
P' = terminal absolute pressure = gage pressure +14.7 lbs. 
For example, 100 cu.ft. of free air, compressed isothermally to 
65 lbs. gage, will occupy a volume: 

100 X 14.7 o (^ 

V =^ — : ^ = 18.4 1; cu.ft. 

65 + 14.7 

Conversely, the volume of free air corresponding to 18.45 
cu.ft. of air at 65 lbs. gage pressure is: 

V'P' 18.45(65 + 14-7) ,, 

V = — p^ = — - = 100 cu.ft. 

P 14.7 

* The deductions of the work formulas used here are given in Chapter III. 

137 



138 COMPRESSED AIR PL.\NT 

By appMng the 8-12^ allowance for losses stated above, 
sufficiently accurate results are obtained for practical purposes. 
As the volumetric output of a given size of cylinder depends 
on the density of the intake air. it will obA-iously be reduced 
when working at an altitude above sea-level (Chap. XIII). 

Second. The size of the compressor may be designated in 
terms of the horse-power developed by the steam end, indicator 
cards being taken while running at normal working speed and 
while the usual volume of air is being compressed. 

Third. The effective horse-power represented by the quan- 
tity of compressed air dehvered is determined from an indicator 
card taken from the air cyhnder. In testing a compressor it is 
customary to take a series of cards, simultaneously from both 
ends of the steam and air cylinders. They may then be com- 
pared, as shown by Fig. 15, Chap. II. 

If indicator cards are not available, the theoretical horse- 
power for single-stage adiabatic compression may be calculated 
by the formula: 



H.P.^ ^^P^'" 



33,0001 ;z — I )L\P 






in which 



P = normal atmospheric pressure, 14.7 lbs. per sq. in.; 

P' = final absolute pressure, lbs. per sq. in.; 

V =the volume of free air compressed per min., cu.ft. ; 

n = exponent of the compression curve. For adiabatic 
compression, 77 = 1.406, and varies down to 1.18 or 1.2, according 
to the efficiency of the coohng arrangements, and whether single 
or stage compression. For the best single-stage compressors, 
;i = 1.3 (approx.). 

For isothermal compression: 



H.P.= -^-» 



OvD 



XPv(xap. log ^) 



Table V shows the horse-powers required, under the condi- 
tions named, to compress one cubic foot of free air per minute: 

* The Xaperian or hN-perbolic logarithm of a number is equal to the common 
logarithm multiplied by the constant 2.302585. 



PERFORMANCE OF AIR COMPRESSORS 



139 



Table V. — Single-Stage Compressors 





Atmospheres 
Absolute, or 

Ratio of 
Compression 
P' 

P' 


Single-Stage Compression, from Atmospheric Pressure 

AT Sea-Level. Initial Temp., 60° F. Horse-power 

Required to Compress i Cu.ft. of Free Air. 


Gage 
Pressure, 


Theoretical Horse-Power. 


Actual Horse-Power (Approx.). 


Lbs. 


Isothermal 
Compression. 


Adiabatic 
Compression. 


Allowance for 
Losses above 
Adiabatic Com- 
pression, 10% 


Allowance for 
Losses above 
Adiabatic Com- 
pression, 15%. 


20 


2.36 


.0551 


.0626 


.0689 


.0720 


25 


2.71 


.0637 


.0741 


,0815 


.0852 


30 


3 04 


.0713 


• 0843 


.0927 


.0970 


35 


3-38 


.0782 


.0941 


•1035 


.1082 


40 


3-72 


.0842 


. 1029 


.1132 


.1183 


45 


4.06 


.0895 


.1115 


. 1226 


.1282 


50 


4.40 


.0950 


. II91 


.1310 


•1370 


55 


4-74 


.0994 


. 1269 


.1396 


. 1460 


60 


508 


. IO41 


•1337 


.1471 


•1537 


65 


5 42 


.1081 


. 1401 


•1541 


. 1610 


70 


576 


.1123 


.1468 


.1615 


. 1690 


75 


6. 10 


. I162 


•1535 


.1688 


•1765 


80 


6.44 


•II95 


•1591 


•1750 


.1830 


85 


6.78 


. 1224 


.1651 


.1816 


. 1900 


90 


7.12 


.1256 


.1703 


•1873 


•1955 


95 


7.46 


.1287 


. 176c 


• 1936 


. 2024 


100 


7.80 


•I315 


.1807 


.1988 


. 2080 


no 


8.48 


.1366 


.1894 


.2083 


. 2180 


125 


950 


.1442 


.2025 


. 2227 


■2328 



In columns 3 and 4 of Table V are the theoretical horse- 
powers required for isothermal and adiabatic compression. 
The results of isothermal compression are wholly unattainable in 
practice, and are placed here only for comparison. The figures 
in column 4 are based on the assumptions that there is no radia- 
tion of heat from the air cylinder, and that the temperature of 
the air after delivery has become normal, its volume being 
therefore reduced to that which is practically available for use. 
These figures include no losses except those due to the heating 
of the air while being compressed. But the full amount of loss 
represented by adiabatic compression can never be suffered 
in the operation of compressors, however imperfect their design. 
The actual compression line is always lower than the adiabatic 



140 COMPRESSED AIR PLAXT 

line, because of the radiation of heat through the cyUnder walls. 
Even in single-stage compressors, properly water- jacketed and 
run at a reasonable piston speed, the compression Kne may fall 
considerably below the adiabatic. Whatever diminution of loss 
is effected by cooling the air in the cyhnder may therefore be 
credited against the other unavoidable losses, partially offsetting 
them, viz.: frictional or mechanical loss in the compressor, 
friction of inlet valves, heating of the intake air by contact 
with the hot metal surfaces, and piston clearance. 

In the absence of indicator cards, estimates based on practice 
may be made of the compressor horse-power. In columns 5 and 
6. of Table \', are shown the actual horse-powers required to com- 
press I cu.ft. of free air, under the conditions stated at top of the 
columns. Thus, in column 5, 10^ is assumed as a fair estimate, 
in case of large, well-designed and operated single-stage com- 
pressors in good running order, of the additional power required, 
over and above that for adiabatic compression. This 10% is 
taken as the algebraic sum of: the loss in purely adiabatic 
compression, minus the effect of ordinary water-jacket cooling, 
plus the four losses mentioned at end of preceding paragraph. 
In column 6, the power consumed in adiabatic compression is 
increased by 15^. which represents relatively poorer work. 
(See also Table VI).* 

It must be remembered that, if the compressor is small, or 
in poor condition, or is run at too high a speed, the required 
horse-power is greater. In such cases, the added percentages 
in column 6 of Table X and columns 5 and 7 of Table VI may be 
increased to 18 or 20%. 

The figures in columns 3 and 4 or 5 and 6 of Table V. fwhich 
are for free air), if multiplied by the corresponding ratios of 
compression (column 2), will give the respective theoretical and 
actual power costs of furnishing i cu.ft. of compressed air, at the 
gage pressures stated. 

* Since the pre\'ious edition of this book was published, the working efficiency 
of the better tjpes of compressor has improved. In view of this, the figures in 
Tables V and Vl. sho\nng the actual horse-power p)er cu.ft. of free air compressed, 
have been materiallv reduced. 



PERFORMANCE OF AIR COMPRESSORS 



141 



Work of Stage Compressors (see Equations 19 and 20, Chap. 
III). The theoretical horse-power for two-stage compression is: 

2X 



H.P.= 



33 



,000 w— iL\P/ J 



For three-stage compression: 

X. X. 3X144 PVwr/P'\^^^ 1 
33,000 n — iL\r/ J 

Reducing the constants, and for a volume of i cu.ft. free air 

r/p/\o.i44 I 
Two-stage, H.P. =0.449 I p- 1 — i 

r/p/\ 0.0962 "I 

Three-Stage, H.P. =0.6735 U^j -ij 
Table VI. — Two- and Three-Stage Compressors 





I 


lORSE-PoWER PER Cu.FT. OF FrEE AIR AT SeA-LeVEL. 


Gage 


Ratio of 
Compression 

_ P' Isot 
p • Co 


hernial 
mpres- 
jion. 


Two-Stage 
Compression. 


Three-Stage 
Compression. 


Pressure, 
Lbs. 


Adiabatic 
Compres- 
sion. 


Actual H.P., 

on basis of 

Adiabatic 

Comp'n 

+ 15%. 


Adiabatic 
Compres- 
sion. 


Actual H.P., 

on basis of 

Adiabatic 

Comp'n 

-1-12%. 


70 


5.76 


II23 


0. 129 


0.148 






80 


6.4 


1 195 


-138 


•159 






90 


7.12 


1256 


.147 


. 169 






100 


7.80 


1315 


•154 


.177 


0.145 


0.162 


120 


9. 16 


1420 


. 169 


.194 


.158 




177 


140 


10.50 


1508 


.181 


.208 


. 169 




189 


160 


11.88 


1583 


.192 


. 221 


.179 




200 


180 


13-24 


1651 


. 202 


.232 


.188 




210 


200 


14.60 


1720 


. 212 


•244 


.196 




219 


250 


18.00 


1853 


.231 


.266 


.213 




238 


30 ) 


21.40 


1963 


.249 


.286 


.22S 




255 


350 


24.80 


2058 


. 264 


•303 


.241 




270 


400 


28.20 


2140 


.277 


.318 


.252 




282 


450 


31.62 


2215 


.289 


•332 


.262 




293 


500 


3501 


2280 






.271 




303 


550 


38.41 


•2339 






.28b 




313 


600 


41 .80 


2393 






.288 




322 


650 


45-21 


2443 






•295 




330 


700 


48.62 


2490 






.301 




337 


800 


55-42 


2574 






-314 




352 



142 



COMPRESSED AIR PLANT 



Tables V and VI may be used for any altitude above sea-level, 
by considering the ratio of compression in column 2. Thus, to 
find the horse-power for two-stage compressionper cu-ft. free air 
at 10,000 ft. altitude and 80 lbs. gage: Ratio of compression 



0.15 0.20 




000S90 



Fig. 81. — Single-stage Compression. 



80+10.07 o / 1 r 1-11 

= ="8.04 (the factor 10.07 bemg the barometric pres- 

10.07 

sure at 10,000 ft., from Table XIII, Chap. XIII). From Table 
VI, column 2, 8.94 compressions (by interpolation) correspond 
to a gage pressure of about 118 lbs. at sea-level, the horse- 
power for which (column 5) equals 0.191 per cu.ft. of free air. 
At sea-level, the horse-power would be 0.159. 



PERFORMANCE OF AIR COMPRESSORS 



143 



Graphic Determination of the horse-power required to com- 
press I cu.ft. of free air to a given pressure (C. W. Crispell, Trans. 
Am. Inst. Min. Engs., June, 191 7). For this the nomograms in 
Figs. 81 and 8ia are more convenient than the formulas. The 
values of P (atmos. pressure) in lbs. per sq. in., for different 



0.15 0,20 




000 30S00 

eo ■* >o <c t- 00 31 o .^ 



Fig. 8 1 a. — Two-stage Compression 



altitudes, are given in Table XIII, colum^n 3 (Chap. XIII). 
Directions for using the charts are printed on them. Note that 
the scale for the H.P. per cu.ft. of air begins at a different 
ordinate for each value of P. By a series of examples in the 
paper quoted, it is shown that the nomograms usually give 
results from 0.7% to 1.6% higher than those from the formulas. 



144 co:mpressed air pl.\nt 

Table M!I is useful for calculations based on volumes and 
mean cylinder pressures. The mean pressures per stroke (col- 
umns 5 and 6) are obtained from the formulas for isothermal and 
adiabatic single-stage compression, by making V=i. thus: 

Mean pressure per stroke ( isothermal > = P X Xap. log ^ 



Mean pressure per stroke 'adiabatic^ =3.463P 



/P'\^-^ 



The work done during one stroke is equal to the mean pressure 
multiplied by the volume in cu.ft. traversed by the piston. 

When air is compressed adiabatically, the relation between 
its temperature T. at the beginning of compression, and the 
terminal temperature T'. is showTi by: 

— ={-r,j . whence T =TItt 

The final temperature may also be found from the formula: 



T' = T. 
1, p 

T and T' being absolute temperatures and P. P' absolute pres- 
sures. 

The compression curve of an air-indicator card may be con- 
structed as follows. P\' being the pressure and volume at one 
point of the curve and F'V the pressure and volume correspond- 
ing to any other point. Designating the index number of the 
curve by .v: 

p7=lv7l . From this, 
'og D^ = •*■ 'og rr ) ; wnence, x = 



'-(v) 



in. * ♦ /; '• 1-406 n — i 1.406— I 
The constant 3.463 = = ; exponent o. 29 = = . 

ti — i .406 n 1-406. 



PERFORMANCE OF AIR COMPRESSORS 



145 



Table VII * 



to 
<u 
l-l 

a 

M 

tc 

i: 

H) 

bo 
ni 

o 


< 

i 

D 
l-l 

D 

Xi 

B 


u 
3 C a^ 


0^ 

1- % 
<-^ 

> 


Mean Pressure per 
Stroke; Air at Con- 
stant Temperature. 
Lbs. 


Mean Pressure per 
Stroke; Air Not 
Cooled. Lbs. 


Temperature of Air; 
JNot Cooled. Deg. 
F. 


o 


I I 


I 










60° 


I 


1.068 


•9363 


9500 


.96 


•975 


■71 


2 


1. 136 


.8803 


.9100 


I 


87 


I 


91 


80.4 


3 


1.204 


-8305 


8760 


2 


72 


2 


80 


88.9 


4 


I. 272 


.7861 


.8400 


3 


•53 


3 


67 


98 


5 


I -340 


7462 


.8100 


4 


■30 


4 


50 


106 


lO 


I 680 


5952 


6900 


7 


.62 


8 


27 


145 


15 


2 . 020 


4950 


6060 


10 


■33 


II 


51 


178 


20 


2.360 


4237 


5430 


12 


.62 


14 


40 


207 


25 


2. 700 


3703 


4940 


14 


•59 


17 


.01 


234 


30 


3.040 


3289 


4538 


16 


-34 


19 


.40 


252 


35 


3-381 


2957 


4200 


17 


-92 


21 


.60 


281 


40 


3.721 


2687 


3930 


19 


-32 


23 


.66 


302 


45 


4.061 


2462 


3700 


20 


57 


25 


-59 


321 


50 


4.401 


2272 


3500 


21 


.69 


27 


-39 


339 


55 


4-741 


2109 


3310 


22 


76 


29 


. II 


357 


60 


5.081 


1968 


3144 


23 


78 


30 


•75 


375 


65 


5-423 


1844 


3010 


24 


75 


32 


-32 


389 


70 


5.762 


1735 


2880 


25 


67 


33 


83 


40s 


75 


6. 102 


1639 


2760 


26 


55 


35 


27 


420 


80 


6.442 


1552 


2670 


27 


38 


36 


64 


432 


85 


6.782 


1474 


2566 


28 


16 


37 


94 


447 


90 


7. 122 


1404 


2480 


28 


89 


39 


18 


459 


95 


7.462 


1340 


2400 


29 


57 


40 


40 


472 


100 


7.802 


1281 


2320 


30 


21; 


41 


60 


485 


105 


• 8.142 


1228 


2254 


30 


81 


42 


78 


496 


no 


8.483 


II78 


2189 


31 


39 


43 


91 


507 


115 


8.823 


II33 


2129 


31 


98 


44 


98 


518 


120 


9.163 


I09I 


2073 


32 


54 


46 


04 


529- 


125 


9-503 


1052 


2020 


33 


07 


47 


06 


540 


130 


9-843 


IOI5 


1969 


33 


57 


48 


10 


550 


135 


10.183 


0981 


1922 


34 


05 


49 • 


10 


560 


140 


10.523 


0950 


1878 


34 


57 


50. 


02 


570 


145 


10.864 


0921 


1837 


35 


09 


SI- 


00 


580 


150 


I I . 204 


0892 


1796 


35 


48 


SI- 


89 


589 


160 


11.880 


0841 


1722 


36 


29 


53- 


65 


607 


170 


12. 560 


0796 


1657 


37 


20 


55- 


39 


624 


180 


13.240 


0755 


1595 


37- 


96 


57- 


01 


640 


190 


13.920 


0718 


1540 


38. 


68 


58. 


57 


657 


200 


14.600 


0685 


1490 


39 


42 


60. 


14 


672 



* Kents' " Mechanical Engineers' Pocket Book," Taken from a table in Rich- 
ards* " Compressed Air," p. 20, 



146 



COMPRESSED AIR PLANT 



The several lines of an air-card have significations entirely 
different from those of a steam card. In Fig. 82 (an ideal card), 
AB is the admission line, BC the compression line, CD the 
delivery or discharge line, and DA the re-expansion line. DA 
represents the effect of the re-expansion of the clearance air, 
on beginning a stroke. Comparing the lines of the air and 
steam cards, they are found to be reversed: 



Air Card. 
Admission line. 
Compression line. 
Delivery line. 
Re-expansion line. 



Steam Card. 
Back-pressure or exhaust line. 
Expansion line. 
Admission line. 
Compression line. 



Delivery Lioe 




Aclmis,sion Line 
Fig. 82 



The elements of an air card, together with the work done, 
as represented by the several lines and areas, are further 
elucidated by Fig. 82a, the compression being adiabatic. 
Let AD = normal atmospheric line at sea-level ; 

AG = P = corresponding atmospheric pressure, acting 
behind the piston at the beginning of the stroke 
(neglecting valve resistances and effect of clear- 
ance of previous stroke) ; 
GE = AD = length of stroke of piston ; 
AB = adiabatic compression curve ; 
BC = delivery line. 



PERFORMANCE OF AIR COMPRESSORS 



147 



At the point B the useful work of compression ceases ; during 
the remainder of the stroke the volume of compressed air V^ at 
the absolute pressure P^ is being forced out of the cylinder 
through the delivery valves. 

The area ABFG = the absolute work of compression. 

The area BCEF = the absolute work of delivery. 

The sum of these areas represents the total absolute work 
(that is, on the basis of absolute pressure) done during compres- 
sion and delivery. 



j^— V'-^- 




T 



hd 



-_i_ 



Area ADEG = work done during the entire stroke by atmos- 
pheric pressure behind the piston. 

Area ABH = net work of compression. 

Area BCDH = net work of delivery. 

Area ABCDA = total net work for entire stroke. 

From this analysis another method may be derived for calcu- 
lating the theoretical horse-power required for compressing air. 
It will be found useful, when a table of temperatures of com- 
pression is available. 



148 COMPRESSED AIR PLANT 

Let 2£' = weight of i cu.ft. of free air = .0765 lb.; 

Cp = specific heat of air at constant pressure = .2375; 
Cp = specific heat of air at constant volume = .1689. 

C 'I'l 

Whence, 7:^ = ;^ = 1.406, and = ^.46^. 

J = Joule's heat unit, taken as 778 ft. -lbs. 
The work represented by the area ABH = 

jxic'XC(r-T)-p(v-vo. 

Also, the work done during dehvery = BCDH; = VXP'-P). 
Hence, the total net work for one stroke of the piston 

= areaABCDA = JX^t'XC(r-T)-(PV-P'VO. 

If Cp be substituted for C, then PV = P'V', according to the 
general equation for air compression, and the total work 

W = JXit'XCp(r-T). 

Substituting for J, w, and Cp, their constant numerical values: 

W=i4.i3(T'-T), 

or, to compress i cu.ft. of air per min., at 60° F., and at sea-level 



H. P. =0.225 



r 

T 



By referring to the last column of Table VH and remember- 
ing that T and T' are absolute temperatures, i.e., thermometric 
temperatures plus 459° F., the horse-power required for com- 
pressing I cu.ft. of free air adiabatically to any gage pressure 
may readily be calculated. 

Other expressions for the mean effective pressures may also 
be deduced from what precedes. M.E.P. for the entire stroke = 



''i^(T-")-J"»KT-''-^-"' 



;)--.] 



M.E.P. during delivery = y(P'-P). 



PERFORMANCE OF AIR COMPRESSORS 149 

The M.E.P. for compression only is found by taking the dif- 
ference between the pressures calculated by the last two formulas. 

The results from the above expressions for work and mean 
effective pressure are theoretical. To find the actual horse- 
power required, allowances must be made for the losses experi- 
enced in the operation of the compressor, as already set forth. 

Compressor Tests. To indicate the observations required to 
secure the data for the complete test of a compressor, together 
with the deductions from the observed data, the following record 
of the test of a compound, two-stage Nordberg compressor, at 
the mines of the Tennessee Copper Co., will be found useful.* 
It will be noted that items 28, 29, and 32 to 35, were necessary in 
this case, because the boiler plant supplied steam for a hoisting 
engine and an independent condenser, as well as for the com- 
pressor. Though the hoist was not running, steam was passing 
continuously to the jackets of the cylinders. The same con- 
ditions would often be met in other tests. The boiler feed water 
was taken from a wooden tank, and during the run this water 
was supplied from two barrels on scales set temporarily over the 
tank. The water of condensation from steam jackets and 
reheater was drawn off continuously and also weighed. The 
calorimeter tests were made with a Peabody throttling calori- 
meter. Eight sets of indicator cards were taken during the 
8-hour test, at hourly intervals. 

Items of Compressor Test (Altitude, 1,800 feet) 

1. Date of test, February 16, 1902 

2. Duration of test, hours 8 

3. Diameter of high-pressure steam cylinder (steam jacketed), inches.. . 14 

4. Diameter of low-pressure steam cylinder (steam jacketed), inches. . 28 

5. Diameter of low-pressure air cylinder, inches 24^ 

6. Diameter of high-pressure air cylinder, inches 15! ' 

7. Stroke of all pistons, inches 42 

8. Diameter of piston rods, inches 2^ 

9. Revolutions of engine, average per minute 90 

10. Piston speed per minute, feet 630 

* Abstracted from an article by J. Parke Channing, Mines and Minerals, May, 
1905, p. 475. 



150 COMPRESSED AIR PLANT 

11. Steam-gage pressure, average, pounds i45 -9 

12. Temperature of steam in steam pipe, average, degrees F 364 

13. Steam pressure in reheating receiver, average, pounds 8 

14. Vacuum in condenser, average, inches 25 . 66 

15. Air pressure in intercooler, average, pounds 22 . 63 

16. Air pressure in receiver, average, pounds 79. 3 

17. Temperature of air at intake, average, degrees F 65 . o 

18. Temperature of air lea\ing low-pressure cylinder, average, degrees F. :2ii . 5 

19. Temperature of air leaving intercooler, average, degrees F 78. 5 

20. Temperature of air leaving high-pressure cylinder, a\-erage, degrees F. 240 . o 

21. Indicated horse-power in high-pressure steam cylinder, average. ... 140. 12 

22. Indicated horse-power in low-pressure steam cylinder, average 15303 

23. Indicated horse-power in both steam cylinders, average 293. 15 

24. Indicated horse-power in low-pressure air cylinder, average 143. 79 

25. Indicated horse-power in high-pressure air cylinder, average 135 . 02 

26. Indicated horse-power in both air cylinders, average 278. 81 

27. Feed water weighed to boilers, pounds 43,343 

28. Re-heater and jacket water from compressor, weighed, pounds 4,081 

29. Average temperature of re-heater and jacket water, degrees F 356. 7 

30. Total heat in i pound of steam at 356.7 degrees F., heat units 1,190. 7 

31. Total heat in i pound of water at 356.7 degrees F., heat units 328.9 

32. Equivalent credit for re-heater and jacket water, pounds 1,127.00 

33. Water weighed from condensation in hoisting-engine jacket, pounds. 1,781 . 00 

34. Steam used to run condenser, pounds 4,300.00 

35. Total credits to feed water, pounds 7,228.00 

36. Total feed water charged to engine, pounds 36,115 .00 

37. Moisture in steam shown by Peabody calorimeter, per cent i .30 

38. Credit for moisture in steam, pounds 473.00 

39. Total steam charged to engine, pounds 35,642 .00 

40. Dry sterm per hour charged to engine, pounds 4,455 .00 

41. Steam consumption per indicated horse-power per hour, pounds. ... i5- 19 

42. Guaranteed steam consumption per indicated horse-power per hour, 

at 92 revolutions per minute, pounds 14. 00 

43. Excess of steam consumption per indicated horse-power per hour 

over guarantee, pounds i . 19 

44. Theoretical delivery of free air per minute at 90 revolutions, cubic 

feet 2,037.8 

45. Slip of air (percentage) 3.0 

46. Actual slip of air per minute, cubic feet 61 . i 

47. Actual delivery of free air per minute, average cubic feet i»976 . 7 

48. Theoretical horse-power required to compress and deli\'er actual 

delivery of air at receiver pressure by adiabatic compression 306.53 

49. Theoretical horse-power required to compress and deliver actual 

delivery of air at receiver pressure by isothermal compression. , 229.00 

50. Actual horse-power shown by air indicator cards 278. 81 

51. Actual horse-power shown by steam indicator cards 293 . 15 

52. Actual horse-power consumed by friction of engine 14-34 

53. Efficiency ratio between steam and air cylinders, per cent 95. i 



PERFORMANCE OF AIR COMPRESSORS 



151 



54. Efficiency ratio between steam and air cylinders guaranteed by 

builder, per cent 87.0 

55. Efficiency of steam, or ratio of steam indicated horse-power to 

theoretical air indicated horse-power, isothermal compression, 

per cent 781 

One of the combined indicator cards, from which the aver- 
ages in items 21-26 were calculated, is shown in Fig. 83. 




Fig. 83. — Combined Cards. Two-Stage Nordberg Compressor. 



In further illustration of the performance of air compressors, 
the combined card from an Ingersoll-Rand " Imperial Type 
10 " two-stage compressor, taken at one of the Berwind- White 
Coal and Coke Company's mines, is given in Fig. 84. 

Figs. 85 and 86 show shop- test cards from the air cylinders 
of an Ingersoll-Rand style '' O " compressor. 

A Record of Field Tests. It would undoubtedly tend to 
secure greater economy in the production of compressed air, if 
superintendents and master mechanics gave more attention to 
the actual results produced by the operation of compressors in 
their charge, and study carefully the frequently unfavorable 
conditions under which these machines are called upon to work. 

Few records of the actual effective horse-power of air com- 
pressors have been published. To express the efhciency, it is 
customary to divide the horse-power of the air cylinder by the 
horse-power of the steam cylinder, as determined by indicator 
cards. The manufacturer of air compressors usually rates his 



152 



COMPRESSED AIR PLAXT 



machine on the basis of its mechanical efficiency, without taking 
into consideration any losses except those of friction. Such a 
criterion does not properiy measure the relative commercial 
values of compressors, nor does it present any indication as to 



V 






■"s Cr&nkEnd 
/ \ 



Scale 20 



y ^^ CrajikEBd 
Scale 16 ^^^ 



L.P. Air Cylinder 



H,P. Air Cylinder 




-8.00- 



7.59- 



1 I 

Fig. 84. — Combined Cards, IngersoU-RHnd, Two-Stage. Direct-Connected, 
Electrically Driven Compressors. Air Cylinders 2^" and i4"X2o". 



Rev. per min 187 

Piston speed, ft. per min 623 

Discharge air pressure, lbs. 93 

Intercooler pressure, lbs 24 

Volumetric efficiency (from card). 95 
I.H.P. of low-pressure cylinder.. . 132 



I.H.P. of high-pressure cylinder. 120.8 

3 Total I.H.P 252.8 

Free air delivered per min. cu.ft. 

(from cardi 1706. 

39c Efficiency compared with adiab'c 97-2% 
Efficiency compared with iso- 
thermal 84% 



the effective horse-power developed under ordinary workmg 
conditions. 

A series of tests were made in 1909 by Richard L. Webb, 
consulting engineer, of Buffalo. X. Y.. on a large number of com- 
pressors in a we.ll-known Canadian mining district. In conduct- 
ing these tests, Mr. Webb had access to plants which had been in 



PERFORMANCE OF AIR COMPRESSORS 



153 



operation for a year or more under normal working conditions, 
and I believe his results will be of value not only to users of air 
compressors, but also to the manufacturers. As a rule, the plants 
tested were in the care of competent machinists and in good 



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Fig. 85.— Card from 3oi"X24" L. P. Air Cylinder of Style " O," Ingersoll-Rand 
Compressor. (St. pressure, 115 lbs.; air pressure, 28 lbs.; r.p.m., 100; spring, 
20.) 

running order, so that the results obtained may be taken as 
representing a fair average of current practice in the United 
States and Canada. The results of a few of these tests are 
given here to show the importance of determining the actual 



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-90 
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Fig, 86.— Card from i8|"X24" H.P. Air Cylinder, Style " O," Ingersoll-Rand 
Compressor. (St. pressure, 115 lbs.; air pressure, 100 lbs.; r.p.m., 100; spring, 
60.) 



efficiency of air compressors when working under the condi- 
tions prevailing in most mines. 

Mode of Conducting the Tests. The following plan was 
employed in each case: First, a boiler test was run for not less 
than two weeks, the coal being carefully weighed, the boiler 
feed water measured, and the total revolutions of the compressor 



154 



COMPRESSED AIR PLANT 



recorded by a revolution counter. From these data, the cost 
per boiler horse-power and the average speed of the compressor 
were determined. Second, the compressor was operated at 
different speeds over its entire range. By means of a meter 
installed in the steam pipe near the throttle, the total steam 
consumed, in pounds per hour, was measured. Indicator cards 
were taken of all cylinders, together with temperatures at the 
air inlet, intercooler, and discharge. To measure the actual 

























































































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Fig. 87. — Compressor Plant No, i. 



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volume of air delivered, a meter was placed in the discharge 
pipe outside of the receiver. A number of simultaneous read- 
ings on all instruments were taken at each speed. From these 
were calculated the total horse-power of the steam and air 
cylinders, the steam consumption, and the total piston displace- 
ment per minute. 

The air and steam meters were of the Dodge type, as modi- 
fied by the General Electric Company, and were operated by 



PERFORMANCE OF AIR COMPRESSORS 



155 



their expert sent for this purpose. The indicators were of the 
Roberts-Thompson and the American-Thompson make, which 
are well known and generally accepted as standard. Their 
springs were calibrated by a standard gage. 



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Fig. 88. — Compressor Plant Na i. 



Results of the Tests. As was to be expected, the friction loss 
was found to be only a small item in the total. The other losses, 
which are frequently overlooked or disregarded, played a large 
part in cutting down the efficiency. The capacity of air com- 
pressors is usually rated according to the volume of the cylinders. 



156 COMPRESSED AIR PLANT 

On this basis, the mechanical efficiency only is given. For 
example, if the horse-power of the air cylinder is loo, and that 
of the steam cylinder no, the efficiency of the compressor is 
rated as 91%. This rating disregards the losses due to adia- 
batic compression, heating of the cylind'er and friction of the 
inlet and delivery valves. The tests show the loss of the engine 
itself to be usually not less than 10% and often considerably 
larger. Losses from the other causes mentioned ranged from 
30% up. 

As IMr. Webb is not at liberty to disclose the identity of 
the particular plants, each test has been designated by a 
number. 

Test of Plant Number One. This consists of three 125 H.P. 
return tubular boilers (one being held in reserve), supplying 
steam for a cross-compound condensing air compressor of 
standard make. The steam cyHnders have Meyer valve gear 
and are 16 in. and 28 in. diameter by 24 in. stroke. The two- 
stage air cyHnders are 28 in. and 18 in. by 24 in. From a two 
weeks' run the following results were obtained. 

Total coal burned, lbs 264,300 

Total feed water, cu.ft 37,459 

Total feed water, lbs " 2,335,568 

Average temperature of feed water, degrees F. i^i 

Average evaporation per lb. coal consumed, lbs 8. 72 

Average revolutions per minute 63 . i 

Indicated horse-power of steam end, corresponding to 63.1 R.P.M. 161 

Corresponding indicated horse-power of air end 123 

Average steam pressure, lbs 115 

Average vacuum, lbs 10. 5 

Average air pressure, lbs q6 

Average temperature of outside air, degrees F : 24 

Average air piston displacement at 70° F., cu.ft 1172 

Average metered output corrected to 70° F., cu.ft 758 

The average evaporation, of 8.72 lbs. of water from 131° F, 
to an average steam pressure of 115 lbs., is equivalent to 9.83 lbs. 
of water evaporated from and at 212° F. per lb. coal consumed. 
At the average compressor speed of 63.1 rev. per min. the 
metered output was equivalent to 758 cu.ft. of free air per 
min., the piston displacement being 1,172 cu.ft. per min. Table 
VIII and Fig. 87 present the principal data of this test run. 



PERFORMANCE OF AIR COMPRESSORS 



157 



To find the average operating results, the curves at 6^ 
revolutions should be followed, at which the indicated horse- 
power of the steam cylinder was 160.8, and that of the air cylin- 
der, 123, showing the mechanical efficiency to have been 76.5%. 



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Fig. 89. — Compressor Plant No. i. 

The theoretical horse-power required to compress isothermally 
I cu.ft. of free air per min. to 96 lbs. (the average pressure) 
is 0.129. The theoretical useful work done by the compressor is, 
therefore, 758 X. 129 or 97.8, and the net total efficiency of the 
compressor is 97.81 -m6i or 60. 



158 



COMPRESSED AIR PLANT 



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160 



COMPRESSED AIR PLANT 



Table IX shows the actual cost of running this compressor 
at ditterent speeds. The data were furnished by the o\\Tier and 
are based on one year's operation. In Fig. 88 these costs are 




w &) so 

R.E.M. 



Fig, 90. — Compressor Plant No. 2. 



platted. sho\snng how the cost per steam horse-power per year is 
affected by the average running speed of the compressor. The 
curve of Fig. 89 shows the operating costs in another way. 
Th^3e costs may be read in terms of 1,000 cu.ft. of free air 



PERFORMANCE OF AIR COMPRESSORS 



161 



compressed to loo lbs. or i,ooo cu.ft. of compressed air at 
loo lbs. gage pressure. 

Test of Plant Number Two. The plant consisted of three 
150 H.P. return tubular boilers, supplying steam for a Corliss 







































































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500 












































































































































































































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" B *' and Labor 

<•<• C " Labor and Int. and Dep. 

" D " " Int. and Dep. and Supplies 




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IDO 



Fig, 91,— Compressor Plant No, 2. 



engine, the air compressor, and steam heating. To determine 
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pipe to the compressor during the test run, so that only the 
portion of steam actually used by the compressor was charged 
to the same. The compressor was duplex, with Meyer valve 



162 



COMPRESSED AIR PLANT 



gear, simple steam cylinders 14 in. by 22 in., and two-stage 
air end, 14 in. and 22 in. by 22 in. stroke, rated by the m.anu- 





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10 20 30 40 50 60 70 80 90 100 

E. P. M. 

Fig. 92. — Compressor Plant No. 2. 

facturer at 1,050 cu.ft. of free air per min. at 105 revs. At this 
plant the test lasted over a month, with the following results: 

Total coal consumed, lbs 459,250 

Total feed water, lbs 2,496,cxx) 

Average evaporation per lb. coal consumed, lbs. . . - 5-46 

Average revolutions per minute 36 . 

Corresponding average indicated horse-power (from curve) 53 



PERFORMANCE OF AIR COMPRESSORS 163 

Hourly readings of the revolution counter were taken, show- 
ing an average speed of 36.05 revs. At this speed the steam 
consumption was 51 lbs. per I.H.P. hour, as measured at the 
throttle, the air meter showing a delivery of 275 cu.ft. of free 
air per min. The total efficiency was 67%. Taking the ording^ry 
method of computing the mechanical efficiency only at the same 
speed, there would be 48 air H.P., divided by 54 steam H.P., 
giving an efficiency of 89%. 

The coal consumption per indicated horse-power per year, as 
shown by the books of the company, amounted at the average 
speed to about 56 tons. Table X, with Figs. 90, 91, and 92, 
present the details of the test on this plant, which was conducted 
in a manner similar to that on plant No. i. 

Test of Plant Number Three. This plant consisted of two 125 
H.P. return tubular boilers, supplying steam for a noncondensing 
cross-compound air compressor of standard make; steam 
cylinders 18 in. and 35 in. by 24 in., air cylinders 14 in. and 
28 in. by 24 in. A two weeks' run gave the following results: 

Total coal burned, lbs 221,190 

Total feed water, cu.ft 34,273 

Total feed water, lbs 2,094,657 

Average temperature feed water, degrees F 154 

Average evaporation per lb. coal consumed, lbs 9-48 

Average boiler horse-power ' 208 

Average revolutions per minute 66 

Average indicated horse-power of steam end, at 66 R.P.M. (from 

curve) 210 

Average indicated horse-power of air end (from curve) 128.5 

Average steam pressure 97 

Average air pressure 97 

Average outside temperature, degrees F 23 

Average air piston displacement at normal speed, cu.ft, at 70° F, . . . 1,372 

Metered output in cu.ft. corrected to 70° F 734 

The average evaporation of 9.48 lbs. of water per lb. of coal, 
from 154° F. to an average steam pressure of 97 lbs., is equivalent 
to 10.4 lbs. of water evaporated from and at 212° F. . At the 
average speed of 66 revs., the displacement was 1,240 cu.ft. of 
free air per min., while the metered output was 734 cu.ft. show- 
ing a net volumetric efficiency of 59%. 



164 



COMPRESSED AIR PLANT 



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PERFORMANCE OF AIR COMPRESSORS 



165 



To determine the conditions in average operation, the curve 
at 66 revs, should be followed (Fig. 93), at which the indicated 
horse-power of the steam cylinders was 210, and that of the air 
cylinders, 128. This shows the efficiency to be 61%, the friction 
loss being 81.5 H.P., or 39% of that delivered by steam end. 
This extremely high friction loss was due to the fact that the 
compressor shaft was out of line, and the plant could not be 

























































































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Fig 93. — Compressor Plant No. 3. 



shut down long enough to rectify it. The details and results 
of this test, given in Table XI and Figs. 93, 94 and 95, are inter- 
esting in exhibiting the inefficiency that may be caused by a 
purely mechanical defect. 

Test Number Four. The results of a test on another plant 
are given in Table XII and Fig. 96, the details of the boiler 
test and of the costs being omitted. In thfs case the compressor 
was of the tandem compound non-condensing type, with Corliss 
valve gear for the steam cylinders. The test shows that, at a low 



166 



COMPRESSED AIR PLANT 



speed, the steam consumption increases more rapidly than with 
the Meyer type of valve. 

Summary. The results of these tests are enlightening, in 
showing the actual an ount of the losses occurring in the com- 



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" B rt and Labor 
" C '» Labor and Depreciation 
" D " " Depreciation 
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R.P.M. 

Fig. 94. — Compressor Plant No. 3. 



90 



100 



pression of air, particularly when the compressor is operating 
under the unfavorable conditions of varying air consumption, 
unavoidable in mining and other work in which machine drills 
play an important part. These losses are always recognized as 



PERFORMANCE OF AIR COMPRESSORS 



167 



existing, by compressor builders and by intelligent users, and 
it is clearly desirable that properly conducted tests should be 
made more frequently. 













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Fig. 95. — Compressor Plant No. 3. 



Again, compressor plants generally develop less power than 
their full rated capacity. An air compressor is essentially a 
variable speed machine, its speed being regulated by some 
form of throttling governor, connected with the air-pressure 
regulator. It is therefore called on to run only as fast as the 



168 



COMPRESSED AIR PLANT 



demand for air may require. It would be well for compressor 
builders to give in their catalogues the horse-power rating at 
different speeds, with a table of efhciencies at different loads 
and speeds, just as is done by some of the manufacturers of 







































































































































































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R. P. M. 

Fig, 96. — Compressor Plant No. 4. 



electrical machinery. Catalogues might also include data re- 
specting the cost per horse-power delivered by the air end of 
the compressor at different working speeds. 



CHAPTER XI 



AIR RECEIVERS 



In its common form the receiver consists of a cylindrical 
shell of steel plate, resembling a steam 
boiler without tubes or flues. It has 
pipe connections to the compressor 
and air main, a pressure gage, safety- 
valve, drain cock, and man-hole. The 
vertical form, Fig. 97, is generally 
preferable, as it occupies less floor 
space. Fig. 98 shows a horizontal 
receiver. The capacity should be 
proportioned to the size of the com- 
pressor. The dimensions range from, 
say, 24 ins., diameter by 4 or 6 ft. 
long, to 48 or 60 ins. by 14, 16, or 18 
ft., the largest sizes having a capacity 
of from 200 to nearly 400 cu.ft. Re- 
ceivers are usually built to stand a test 
of 165 lbs. cold-water pressure, for 
working under pressure of 100-120 lbs., 
higher pressures than this being rarely 
necessary for mine service. The shells 
are single-riveted on circular seams 
and, except for small sizes, double- 
riveted on longitudinal seams; the 
heads being dished or hemispherical. 
For best results, the receiver should be 
placed close to the compressor, or in any 
case not more than 40-50 ft. distant. 

The principal functions of an air-receiver are: (i) to eliminate 
the pulsating effect of the strokes of the compressor piston and 

169 




Fig. 97. — Norwalk Vertical 
Receiver. 



170 



COMPRESSED AIR PLANT 



prevent rapid fluctuations of pressure; (2) to minimize the 
frictional loss attending the flow of air through the hnes of 
piping; (3) to serve in some degree as an equahzer and reservoir 
of power; (4) to cool the air before it passes into the mam, thus 
causing it to deposit a part of its moisture in the receiver, whence 
it is drained off. 

Regarding the first point, the volume of the receiver should be 
sufficiently great in proportion to that of the compressor cylinder 
to prevent any material rise of receiver pressure by the volume 
of air forced into it at each stroke. If the compressed air 
passed directly into the main, large fluctuations of pressure 
would occur, accompanied by periodic acceleration of flow. 



12 AIR OUTLET -j . 1 P0R4' 

J J SAFETY VALVE 




Fig. 98. — Horizontal Receiver-Aftercooler (IngersoU-Rand Co.). 



This would increase the frictional resistance in the pipe, and 
at the end of each stroke the compressor piston would have to 
force the air out of the cylinder against a pressure momentarily 
greater than the normal. The violence of the discharge pulsa- 
tions is obviously greater in a single cylinder than a stage com- 
pressor, because the total discharge must take place from a 
cylinder of larger diameter in a smaller proportion of the length 
of stroke than is the case \^ith the high-pressure cyKnder of a 
stage compressor. In the latter the delivery valves open 
earher in the stroke, and the air pipe is about ©ne-half the 
diameter of the cylinder. 

The second function of the receiver is best fulfilled by placing 
an auxiUary receiver near the point where the compressed air is 



AIR RECEIVERS 171 

used. Just as the receiver at the compressor diminishes the 
momentary rise of pressure in the main due to each stroke of the 
piston, so a second receiver close to the machine using the air 
prevents a drop of pressure as each cylinderful of air is drawn 
off. By reducing the fluctuations of pressure the two receivers 
maintain a practically constant flow of air through the main 
connecting them, thus minimizing friction and loss of pressure.* 
For mine service the second receiver would be placed somewhere 
underground; always an advantageous arrangement when the 
air main is long. Underground receivers are not often used 
for air drills alone, but they are a necessity for pumps and hoists 
run by compressed air. They are also useful in permitting a 
further deposition of moisture, thus rendering the air dryer 
and more suitable for expansive-working engines. To accom- 
plish this most effectually, the underground receiver should be 
placed at the point in the pipe line where the air' has reached 
its lowest temperature. 

Underground receivers are usually made like those installed 
near the compressor. Sometimes, however, a chamber is 
excavated in the rock, and the walls cemented or asphalted 
tight. The chamber is closed by a brick dam of two parallel 
walls, with a 2-in. layer of cement between them. In the dam 
are set a cast-iron man-hole with suitable cover, the pipes for 
connecting ^Vith mains to the different working places, and 
a drain pipe and cock close to the floor. The latter is opened 
from time to time, to blow out the accumulated water and 
sediment. A pressure gage is attached to the man-hole cover. 
Such reservoirs may be built to cost much less (for large sizes) 
than ordinary shell receivers of equal capacity. 

The third function of the receiver is apt to be exaggerated. 
While it acts to a limited extent as a reservoir of power; yet, 
to be of much practical use in this respect, it must be very large. 
For example, take a 20-in. compressor, working at 60 lbs. pres- 
sure. To meet the demand for only i minute after the com- 
pressor is stopped, and not have the pressure fall more than 

* Questions relating to the flow of air in pipes, and frictional losses are 
discussed in Chap. XVL 



172 COMPRESSED AIR PLANT 

15 lbs., the receiver \Yould have to be 5 ft. diameter by 50 ft. 
long. Again, if the compressor were running at a constant speed 
and the demand for air should suddenly increase 259c5 as might 
happen in starting several more machine drills, a receiver of the 
size mentioned could meet the extra demand only 4 minutes. 
Long pipes of large diameter assist in equahzing the flow of air, 
but their use does not preclude the necessity of receivers. It 
is much cheaper to employ piping of moderate size, in connection 
with a receiver of generous dimensions. 

The fourth function of the receiver is probably the most 
important. Considerable moisture is always present in com- 
pressed air, due to the natural humidity of the atmosphere, 
especially in warm weather. The velocity of the air coming 
from the compressor is greatly reduced on entering the receiver; 
and on cooling the air deposits part of its moisture, which 
otherwise would be conveyed into the piping, and thence to 
the machines using the air. Moisture tends to wash away the 
lubricant, and so increase wear, and consequent leakage of air 
and loss of economy. This is especially true of high-speed ma- 
chines, as drills and small air hoists, in which the wearing sur- 
faces are limited in area. Moisture collecting in pipe lines also 
causes " water hammer," reduces the air passage by accumulat- 
ing at low points, and in winter may freeze and burst the pipes. 
Wet air, freezing in drills, etc., may clog the exhaust and increase 
back pressure. Moreover, hot air in pipe lines causes expansion, 
and when the pipe cools during a shut-down, contraction takes 
place, all of which tends to leaky joints. The receiver should 
be large enough to drain the air thoroughly. In the ordinary 
sizes of receiver the results are usually quite imperfect, because 
the air passes too rapidly to permit a large drop in temperature. 
The inlet and outlet pipes of the receiver should be placed in 
proper relative positions. If at opposite ends, and especially 
if the pipes point toward each other, a strong through current 
is caused, and the air passes out without ha\ing had time to 
cool or to drop much of its entrained moisture. One mode of 
arranging the pipe connections is to place the inlet on one side, 
near the end of the receiver, while the outlet is at the opposite 



AIR RECEIVERS 



173 



end, in the middle of the head. The air is thus forced to change 
its direction of flow. Or, as in Fig. 97, both pipes may be 
connected near the top, the outlet pipe being carried nearly 
to the bottom, where the air is likely to be slightly cooler (and 
dryer) than at the top. As the inlet pipe shown in this , case 
is connected tangentially to the 
periphery of the receiver, a 
rotary motion is imparted to the 
body of air, so that each particle 
remains longer in the receiver 
and under its cooling influence. 
Some receivers have baffle-plates 
for the same purpose, as in Fig. 
98. Part of the lubricating oil 
carried over from the compressor 
cylinder is also deposited in the 
receiver. At intervals, accord- 
ing to atmospheric and other 
conditions, the water and oil 
are drained off. 

Receiver Aftercoolers. A re- 
ceiver of ample size, placed close 
to the compressor, tends to 
economize power; because,what- 
ever cooling is accomplished re- 
duces the temporary increase of 
pressure due to the heat of 
compression. Hence, in forcing 
the air out of the cylinder 
against the receiver pressure, the 
piston consumes less power 
than if the air were left to cool 
gradually in a long length of piping. As the heat of compression 
is always lost before the air is used, this saving is worth while, 
however small it may be, since it is produced without cost. 
This consideration has of late led to, the employment of " receiver 
aftercoolers," Fig. 98 shows one form, somewhat similar to the 




Fig. 99. — Ingersoli-Rand Vertical 
Aftercooler, Type " VK." 



174 COMPRESSED AIR PLANT 

intercooler in Fig. 49.* Fig. 99 shows a recent design, made 
in 6 sizes, from 2o| ins. diameter by io| ft. long, to 45 J ins. by 
19 ft.; cooling surface, 152-2012 sq. ft. Horizontal coolers are 
also furnished. 

These af tercoolers cool and dry the air more thoroughly than 
ordinary non- tubular receivers, and so minimize the troubles 
referred to on p. 172. In Fig. 99 the shell is steel, with cast 
heads. The arrows show the direction of flow of air and water. 
By means of the open funnel in the water discharge pipe, the 
flow of water is seen, and regulated as necessary. A plate in 
front of the air discharge prevents escape of water deposited 
by the air. Such an aftercooler will reduce the temperature 
of compressed air to within 15° or 20° of that of the entering 
water. The Ingersoll-Rand Co. gives the following figures for 
air at 80-100 lbs., from a two-stage compressor: 

Temp. Cooling fo^o'cuTt' f^ree a!r 

Water. ^''' p"er min ' ^'' 

50° F 120 

60° F 1 50 

70° F 180 

80° F 210 

Capacity of Receivers. No exact rule can be given. Re- 
ferring to the statements on pp. 1 71-172, regarding the third 
function of receivers, the capacity should be sufficient to pre- 
vent rapid or great fluctuations of air pressure, and must 
therefore depend largely on the kind of service and local con- 
ditions. 

A safe rule for ordinary rock-drill service is to allow a 
receiver capacity of 100 cu.ft. per 800 to 1000 cu.ft. of free 
air compressed per minute For stationary, constant-running 
engines, like pumps, the capacity may be smaller. 

* See an article by Frank Richards, in Compressed Air , Jan., 1907, p. 4329. 



CHAPTER XII 
SPEED AND PRESSURE REGULATORS FOR COMPRESSORS 

If the air consumption were constant, no more regulation 
of the compressor's speed and power would be required than 
that furnished by the steam governor, to take care of fluctuations 
in boiler pressure or accident to the mechanism. But there are 
usually wide variations in the rate at which the air is used. 
In event of a sudden decrease in consumption, the compressor 
must be slowed down, or air will be blown off at the receiver 
safety valve. Since a cubic foot of compressed air costs more 
than a cubic foot of steam, the compressor must have some device 
for coordinating the quantity of steam admitted to the steam 
end with the variable receiver pressure, thereby regulating the 
piston speed in accordance with the demands upon the air end. 
Furthermore, it i^ not enough to provide only for varying the 
speed of the compressor. At times, the consumption of air may 
cease entirely for a short period, and, to avoid bringing the com- 
pressor to a standstill, provision should be made for unloading 
the air end. Useful work then stops for the time being, the 
compressor consuming only enough steam to turn its centers. 

Numerous regulating and unloading mechanisms have been 
devised, so that instead of requiring the constant attendance 
of an engineer, the compressor operates automatically under 
wide variations of load. These devices may be classified under 
two heads: (i) speed governors and pressure regulators; (2) un- 
loaders for the air cylinders. 

Speed Governors and Pressure Regulators. Speed gover- 
nors, usually of the centrifugal or flyball type, may be applied 
•to the steam cylinder merely to regulate speed; or their action 
may be controlled by the receiver pressure so as to regulate 
both speed and pressure. The air cylinder is not completely 

175 



176 



COMPRESSED AIR PL.\XT 



unloaded at any time, the compressor being simply speeded 
up or slowed down according to the rate at which the air is used. 
The flyball governor of the regulator tx-pe is illustrated 
by Fig. ICO. The stem // of the throttle valve connects with 
the spindle of the ball governor, by which the speed of the com- 
pressor is controlled. At p is the bevel gearing for dri\'ing the 
governor, a small pulley being mounted on the gear shaft and 




Fig. ioo. — Clayton Go\emor and Pressure Regulator. 

belted to the crank-shaft of the compressor. The action of the 
ball governor is modified by the weighted lever / and the air 
cvlinder 7, which is connected to the air receiver by the pipe k. 
When the receiver pressure exceeds its assigned limit, it raises 
the piston and weight, and shuts off steam by forcing down the 
throttle valve //, the pressure of the lever being appHed at /. 
The governor is adjusted to its work by the spring and thumb- 
screw m. actino: on the lever n, which tends to keep open the 



SPEED AND PRESSURE REGULATORS 



177 



throttle against the downward pressure of the lever i upon the 
valve stem. The spring o eases the drop of the weight when 
the air pressure falls. 

Flyball governors of several types are used on the IngersoU- 
Rand, Sullivan, American and other compressors. 




Fig. ioi. — Norwalk Pressure Regulator. 

The design of the Norwalk governor (Fig. loi) is entirely 
different. Above a balanced throttle valve in the main steam 
pipe is set a small air cyhnder B. At the side of the cylinder 
is a spring-controlled valve, connected by a pipe with the receiver, 
or with the air main leading to it. The spring of this valve is 
adjusted so that the air will lift the valve, and pass through it, 
at any desired pressure. When the receiver pressure exceeds 



178 COMPRESSED AIR PLANT 

this limit the valve allows air to pass under the heavy piston 
in the cyHnder B, raising it and partly closing the throttle. 
If no escape were provided the piston would be forced at once 
to the top of the cylinder. To regulate its movement and 
prevent shutting off the steam completely, a tapered recess is 
cut in the piston rod of this cylinder, at the point where it passes 
through the lower-head (indicated at R, in the small cut to 
left of main figure). As the piston is forced upward by the air 
pressure the area of the opening formed by the slotted stem 
furnishes a graduated escape for the air, and so regulates the 
small piston's movement and through it the position of the 
throttle valve. The upward movement of the piston is still 
further regulated by the screw stop and spring in the top of 
the cylinder. This can be so adjusted that, when the piston 
reaches its highest point, the throttle valve still admits enough 
steam to keep the compressor turning its centers. 

With governors of the preceding t}'pes, the operation and 
control of the compressor is not automatic under all conditions, 
but they answer the purpose for some kinds of service. In case 
no air is drawn from the receiver, the compressor is brought 
nearly to a standstill; then, if the pressure continues to rise, 
a Httle air will blow off at the receiver safety-valve, or the 
compressor may be stopped by closing the throttle. 

A combined governor and pressure regulator, with unloading 
attachment, as employed by the Sullivan ^Machinery Co. for 
steam-driven com.pressors, illustrates a mode of control that has 
been adopted by several builders, though with many variations 
in details (Fig. 102). The split-ball governor (11), belt-driven 
from the crank-shaft to the pulley (20), accompanied by the 
tightener (43), controls the steam throttle (3). Comiected 
with the governor spindle and throttle valve stem, at (28), is a 
lever (25), the position of which is influenced by the centripetal 
action of the set of springs (31, 32, and 26). By screwing up 
or down the hand-wheel and speeder screw (5), this system of 
springs (and with them the governor) is set to run the compressor 
at any desired speed. The other element of the governor is the 
air-pressure device, which, by the position of the plunger in the 



SPEED AND PRESSURE REGULATORS 



179 



small air cylinder (i8), brings the springs into action in the 
order of their strength, thus producing movement of the lever 
(25). The pressure device is connected to the receiver by the 
union valve (33), admitting air to the little cyHnder (27), the 
piston of which operates a needle valve. This valve is ,held 
closed against any desired minimum air pressure by the adjust- 
able weight (36) and the regulating screw and spring (37 and 3.8). 



Oil Gov, Bere 




Fig. 102. — Sullivan Governor and Unloader. 



When the receiver pressure exceeds the normal, it opens the 
needle valve and admits receiver air to the cylinder (18). As the 
pressure increases, the plunger in (18) rises against the counter- 
spring (26) and through the lever (25) tends to close the main 
steam throttle (3), thus slowing the compressor. Total stoppage 
is prevented by screwing down the nut of the stop-screw (23), so 
as to limit the upward movement of the plunger (18), which acts 
intensively, being so proportioned that a variation of only 2 or 3 



180 COMPRESSED AIR PLANT 

lbs. receiver pressure is multiplied to say 40 lbs. in its action 
on the governor. A sensitive control is thus produced within 
narrow limits of working pressure. To prevent \dolent move- 
ments of the pressure element, in case of sudden changes of 
receiver pressure, the plunger in (18) has an oil dash-pot. 

A somewhat similar pressure regulator and unloader is used 
on the Franklin compressor.* 

For steam-driven compressors of the Corliss t}^e, as built by 
the IngersoU-Rand, Xordberg, Laidlaw-Dunn-Gordon, SuUivan, 
Allis- Chalmers, and some other companies, the regulators act 
in conjunction wdth ball or other centrifugal governors. They 
control by changing the point of cutoff in the steam cylinder. 

The Laidlaw-Dunn-Gordon governor (Fig. 103) is an example. 
Air from the receiver enters the small cylinder A, the piston 
of which is weighted. The action of the lever B is adjusted by 
the coil spring C. This lever is linked to a floating lever D, 
pinned to the vertical side rods of the ball governor. D is con- 
nected by the link E to the bell-crank F, the lower arm of which 
is connected through the long horizontal rod G to the Corliss 
steam gear. By this system of levers, the movement of G, and 
through it the point of cutoff, is under the combined control 
of both ball governor and of the receiver pressure, as influencing 
the position of the piston of the cylinder A. The arm H is 
pivoted at the foot of the governor post. Connected to it are 
the cam I and the idler pulley J, which rests on the governor 
belt. In case the belt breaks, the idler pulley falls and the cam 
allows the governor to drop, thus shutting oft' steam and pre- 
venting the compressor from racing. 

One of the IngersoU-Rand regulators, used for compressors 
with compound steam ends, controls speed by varying the cut- 
off of the high-pressure cylinder (Figs. 104, 105). This governor 
contains an oil pump a, chain-driven from the compressor crank 
shaft to sprocket s. The oil from the pump enters the plunger 
chamber c under pressure, and acts to force upwards the 
plunger d, which carries weight e. Vertical movement of d is 
transmitted by rack g, pinion //, and sprocket i, through a 

* Mines and Minerals, May, 1905, p. 504. 



SPEED AND PRESSURE REGULATORS 



181 



chain, to the cutoff valve stem, causing the stem to rotate and 
thus vary the cutoff. Admission of oil to c is controlled by a 
by-pass valve/. This valve is operated through the panto- 




FiG. 103. — Laidlaw-Dunn-Gordon Air Governor. 



graph motion 7, by the movement of either the plunger J, or 
the air regulator weight and lever k. 

When the compressor speed (and therefore the speed of the 



182 COMPkL^-ll .Uv PLAXT 

oil pump) i^xxlaces sufficient oH pre^ure to force phmger d 
Inward, valvf / (^)0is and checks the rise of d until the speed 
increases sti'J' ^Jrther. The maximum and minnnum speed 
scre-"^ " .'e set at the factor\" to limit the maximum 

niTt^ r : -- ^ 'event sucb a short enins^ of the cutoff as 

A„--Cy':- ie: Uniiaders. Thesr txeriii-r complete control 
when the compressor is belt-driven: and also for steam-driven 




FKi- iC4_ — IngasoB-Raiwi " XFV " Antomatic Air Gc»srer!y>r- 

compressoi^ when used in conjunction with a governor. In 
steam compressors, as the consumption of air decreases the 
throttle is first nearly dosed; then, if it ceases altogether, 
the unloading mechanism either shuts off the intake air or 
hc^ds open the discharge \~alves. thus admitting air at receher 
pressure to both aids of the c\-liiider. In either case the pres- 
sures on (^ipc^te ddes of the piston are balanced and useful 
wmk ceases, though the compressor continues to run slowly. 
The Rand " Impmal " unloader. for compressors driven by a 
belt or direct-connected electric motor, is an example of this 



SPEED AND PRESSURE REGULATORS 



183 




o 

e 

> 

o 

O 



B 

o 
■*-> 



> 



c 



be 



<U 

Q 



O 



o 



184 



COMPRESSED AIR PLAXT 



t}pe of regulator. It is placed in the intake pipe, and shuts 
oti the air from the inlet valves when the receiver pressure 
rises above the set Hmit. In Fig. io6 the the intake air enters 
as shown by the arrows. The smaU chamber (60) is connected 




SECTIONAL VIEW 
RAND IMPERIAL UNLOADER 



K Pipe 
from Receiver 



Fig. 106. 



by a i-in. pipe TNdth the receiver. As the pressure increases, 
the piston (57) moves against the spring (56), admitting receiver 
air through the small ports on the left of the piston to the lower 
side of the plunger valve (61). On reaching its seat this plunger 
closes the air intake. The spring (56) may be adjusted by the 



SPEED AND PRESSURE REGULATORS 185 

screw-plug (55) for any required working pressure. As the 
receiver pressure falls, on increased consumption of air, the spring 
forces down the piston (57). This closes the lower small air port, 
leading to the under side of the plunger valve (61), and opens 
the upper horizontal port, connecting with the open screw-plug 
(55). The air below the plunger valve is thus exhausted, 
causing the latter to reopen the intake passage. The com- 
pressor then resumes useful work. An unloader similar to the 
above is used in some of the AlUs-Chalmers compressors. An 
automatic " choking " controller is shown in Fig. 29, Chap. II. 
The Ingersoll-Rand Co. also makes a clearance controller, for 
unloading the air end of small power-driven compressors. It 
varies the clearance volume of the cylinder by cutting in or 
out some of the discharge valves. A small air cylinder, con- 
nected with the receiver, is attached to the compressing cylin- 
der. As the receiver pressure increases, the weighted piston 
of the controller cylinder rises higher. Inserted in the side 
of this cylinder is a series of small pipes, each connected by 
branches with a discharge valve on each end of the main cylinder. 
These valves are thus released from the receiver piessure suc- 
cessively, as the pressure increases, and the work done by the 
compressor is proportionately reduced. When normal receiver 
•pressure is restored, the valves close automatically, and com- 
pression and delivery are resumed. This controller has the 
disadvantage of suddenly releasing and resuming the load. 

Another type of unloader is employed on the Nordberg 
constant-speed, variable-delivery compressor. It is for motor- 
driven machines, with Corliss air valves, and operates by closing 
the inlet valve before the forward stroke is completed. During 
the remainder of the stroke, the air already admitted to the 
cyhnder expands below atmospheric pressure, and is then com- 
pressed on the return stroke. This is practically equivalent to 
varying the working length of stroke. 

The valve gear of this compressor is shown in Figs. 107 and 
108. In Fig. 107 the wrist-plate w is driven by the rod a from 
an eccentric on the fly-wheel shaft; another eccentric operates 
the releasing mechanism through the rod b, which oscillates 



186 



COMPRESSED AIR PLANT 




a 



o 
U 



< 



Q 
*c 

> 

<u 
<u 
D- 

■<-> 
en 
C 
O 

U 



O 



O 



SPEED AND PRESSURE REGULATORS 



187 



the arm c about the fixed center d. Swivelled to the lower 
end of c is a 3 -armed rocker. The arm i is linked by the rod j 
to the radius fork k, which in turn is connected to the pressure 
governor /. The arms g and h of the rocker, through the rods e 




Fig. 108. — Detail of Valve Gear Shown in Fig. 107. 



and /, operate the knock-off or releasing cams n and 0, attached 
to the inlet-valve spindles. When the compressor is working 
regularly, under normal air consumption, the rocker arms g and 
h remain vertical, under the action of the eccentric rod h, and 



188 COMPRESSED .\IR PLANT 

impart equal movement to both knock-off cams. If. however, 
the receiver pressure increases, the rocker arm / moves upward, 
the arms g -and // take an incHned position and, through the 
rods e and/, the point of release of the valves is altered. 

The releasing mechanism is shown by Fig. io8. Mounted 
on the valve spindle is a rocker ha\'ing three arms. a. h, and c. 
The -^Tist-plate link is connected to arm j. the releasing latch 
d to arm h, and the governor cam-arm e to arm c. By the rod/, 
€ is connected also to the governor as explained above, and 
hence has a compound motion; it swings bodily above its 
swivel pin at the top. and its position is adjusted laterally by 
the action of the governor. The cam slot has two circular arcs, 
struck from the center at the upper end of e. vr^th. an inclined 
jog connecting them. Since the roller on the arm g s\\ings 
about its center under the action of the cam groove, as the cam 
is moved from the main eccentric by the rod /'. the latch d is 
alternately released and engaged, when the roller passes the 
jog in the cam. The point of the stroke at which release takes 
place is determined by the governor. 

Figs. 109. no and in. are a set of indicator cards from 
a two-stage compressor pro^'ided with this regulating mechan- 
ism, and running at 74 revs, per min. The upper card in each 
cut is from the intake cylinder, the lower from the high-pressure 
cylinder. Fig. 109 shows the cards when working at nearly full 
load. Fig. no (half load) illustrates the action of the regulating 
gear. Taking the crank-end card C, the inlet valve remains 
open from the beginning of the stroke, at a. approximately to 
mid-stroke h, at which point the releasing gear acts and the valve 
closes. From h to the end of the stroke, at r, the air in the 
cylinder expands below atmospheric pressure. On the return 
stroke, the compression line nearly coincides vdxh the expansion 
line from c, until atmospheric pressure is reached at the point 
b. after which compression proceeds as usual. The action of 
the inlet valves of the high-pressure cylinder is the same, except 
that the expansion and re-compression of the air is from receiver 
pressure, instead of atmospheric pressure. In Fig. in the 
cards show the small amount of work done when the compressor 



Speed and pressure regulators 



1^9 




T3 
O 






O 

l-H 



ri 
O 









T3 

O 



o 



o 

I— I 



190 



COMPRESSED AIR PLANT 



is under nearly zero load. To simplify the mechanism, each 
cylinder has its own governor. 

The Ingersoll-Rand " RA-39 " controller (Fig. 112) is 
another device for cutting off air at the intake. It consists of a 
balanced disk valve a, inserted in the intake pipe and held 
open by the spring g. When the receiver air, entering at e, 
exceeds the desired pressure, it forces the diaphragm c to the 
right against the spring /, the resistance of which is adjusted 
by the screw behind it. Attached to the diaphragm is a needle 
valve b, which admits receiver air against the hollow-piston d, 



Oil Cnp 




Fig. 112. — Ingersoll-Rand "RA-39" Controller. 



thus closing the valve a. When the receiver pressure falls, 
the needle valve closes, the air leaks out from behind the 
piston d, and the regulator vdAve is forced open by spring g, 
again admitting air to the compressor. This controller is 
applicable to belt- or motor-driven compressors. It may be 
used in connection with the fly-wheel governor for steam- 
driven compressors. When a variable speed is desired, a fixed 
cutoff with a variable speed throttling governor is substituted. 



CHAPTER XIII 
AIR COMPRESSION AT ALTITUDES ABOVE SEA-LEVEL 

Because of the diminished density of the atmosphere, air 
compressors do not produce the same resuhs at high altitudes 
as at sea-level. Their effective capacity is reduced because a 
smaller weight of air is taken into the cylinder at each stroke. 
It is necessary, therefore, to modify the figures relating to the 
capacity and performance of compressors, as set forth in the 
first part of Chap. X. This matter is of especial importance 
in connection with mining operations, because of the large num- 
ber of mines situated in elevated mountain regions. The rated 
capacities of compressors, in cubic feet of air, as given in the 
makers' catalogues, are for work at normal atmospheric pressure, 
and due allowance must be made for decreased output at eleva- 
tions above sea-level. This reduction in output, which is usually 
also tabulated in handbooks and catalogues, should receive 
due consideration in order to avoid serious errors. For example, 
the volume of compressed air delivered at 60 lbs. pressure, at 
10,000 ft. elevation, is only 72.7% of the volume delivered 
at the same pressure by the same compressor, at sea-level. 
In other words, a compressor which at sea-level will supply power 
for 10 rock-drills, will at an elevation of 10,000 ft. furnish air 
for only 7 drills. 

The foregoing statement relates only to the volumetric 
capacity of the compressor. It must be remembered that the 
heat of compression increases with the ratio of the final absolute 
pressure to the initial absolute pressure. As this ratio increases 
with the altitude, more heat will be generated by compression 
to a given pressure at high altitudes than at sea-level. This 
additional heat temporarily increases the pressure of the air 

191 



192 



COMPRESSED AIR PLANT 



in the cylinder, while under compression, and more power is 
therefore required to compress and deliver a given quantity of 
air. The corresponding loss of work, due to the subsequent 
cooling of the air in receiver and piping, also increases with the 
altitude. 

Contrary to a common impression, the volume of air delivered 
by a given compressor does not bear a constant ratio to the 
barometric pressure, but at different altitudes this volume 
decreases slower than the barometric pressure. This relation 
may be shown as follows:* Two ideal indicator cards are 




Fig. 113. 

represented in Fig. 113, one of a compressor working at sea-level, 
with an initial pressure Pi, the other at an altitude with a lower 
initial pressure P2. The initial volume V and the final gage 
pressure P are the same for both compressors, P3 and P4 being 
the respective final absolute pressures. Vi and V2 are the 
final volumes, corresponding to the dotted isothermal curves, 
these volumes being taken as the basis, because they are those 



* The general method of demonstration here given, together with Fig. 113 and 
accompanying table, are taken by permission from an article by F. A. Halsey, in 

American Machinist, June 2, 1898, p. 27, 



AIR COMPRESSION AT ALTITUDES ABOVE SEA-LEVEL 193 

to which the compressed air will eventually shrink on losing 
the heat of compression. From the theory of air compression, 



and 



VPi==ViP3,or^=|^, (i) 

VP2 = V2P4,or^=|^ (2) 



But since P3 = Pi +P, and P4 = P2+P, equations (i) and 
(2) may be written: 

V Pi+P P . 

and 

V2" P2 "'■^P2 ^"^^ 

Dividing equation (3) by equation (4) : 

y- = ' — ^, or Vi :V2 : : i+p- : i+p-. . . (5) 
^+P^ 

This gives an expression for the ratio between pressure and 
volume at sea-level and for any altitude above sea-level, of which 
the corresponding barometric pressure is P2. Thus, let P2 = 
10 lbs., P = 9o lbs., and Vi (from Table VII, p. 145) =0.1404 cu.ft. 
By substituting these quantities in equation (5), V2 is found 
to be 0.0999, or approximately o.i cu.ft. 

In Table XIII, columns 4 and 5, are given the relative 
volumetric outputs, at gage pressures of 70 and 90 lbs., of 
a compressor working at different altitudes, the figures being 
percentages of the normal output at sea-level. These per- 
centages have been derived by Mr. Halsey from equation (5), a 
constant loss of initial pressure of 0.75 lb. being assumed, to 
allow for the resistance presented by the inlet valves (see 
Chap. VII) ; that is, for practical purposes the sea-level atmos- 
pheric pressure is taken as 14, instead of 14.7 lbs. The figures 
in columns 4 and 5, which are for the ordinary range of pressure 



194 



COMPRESSED AIR PLANT 



employed in mining, show that, though there is a difference of 
20 lbs. between the two gage pressures, the outputs vary only by a 
few thousandths and may often be neglected.* Wide differences, 
however, occur in the other columns. The method of computing 
compressor horse-power for a given number of machine drills, 
working at altitudes above sea-level, is given in Chap. XX, p. 298. 

Table XIII 



-l-> 

fe 


Barometric 
Pressure. 


Relative 


Out- 


M.E. 
Gage P 


P. for 


c 

D 


u.ft. Piston 
isolacement 


Cu.ft 

press* 

per I;I 

Gage P 


Com- 
;d Air 


JJ 


Inches 
Mercury. 


Lbs. per 
Sq.in. 


put for Gage 
Pressure. 


ressure. 


per I.H.P. foi- 
Gage Pressure. 


I. P. for 

ressure. 


Alii 


70 lbs. ' 90 lbs. 


70 lbs. 


go lbs. 


70 lbs. 


90 lbs. 


70 lbs. 


90 lbs. 


I 


2 


3 


4 j 5 


6 


7 


8 


9 


10 


II 





30.00 


14- 75 


1 .000 ii .000 


33 I 


38.2 


6.93 


5-99 


I. 144 


.801 


I, ODD 


28.88 


14.20 


.967 




966 


32.6 


37 


6 


7 


03 


6. 09 


I. 123 


-787 


2, ODD 


27.80 


13.67 


•935 




933 


31 I 


36 


9 


7 


15 


6. 20 


1. 103 


•773 


3. ODD 


26.76 


13-16 


.904 




900 


31-5 


36 


3 


7 


27 


6.31 


I.0S4 


•759 


4.033 


25 76 


12.67 


.873 




869 


31.0 


35 


6 


7 


39 


6.43 


1.065 


.746 


5, ODD 


24- 79 


12. 20 


.843 




839 


30- 5 


35 





7 


51 


6.55 


1.046 


■733 


6, ODD 


23.86 


11.73 


•813 




809 


30.0 


34 


3 


7 


65 


6.67 


1.028 


.720 


7. ODD 


22.97 


11.30 


•785 




780 


29.4 


33 


7 


7 


80 6. 79 


I .Oil 


.708 


8,O0D 


22 . II 


10.87 


• 758 




751 


28.9 


33 


I 


7 


94 


6.92 


•994 


•695 


9.OOD 


21 . 29 


10.46 


.731 




723 


28.3 


32 


5 


8 


09 


7.06 


.976 


• 683 


lO.ODD 


20.49 


10.07 


■705 




696 


27.8 


31 


8 


8 


24 


7.20 


•959 


. 670 


1 1 .000 


19.72 


9.7c 


.680 




671 


27.4 


31 


2 


8 


39 


7-34 


.942 


.658 


I 2. ODD 


18.98 


9-3< 


.656 




647 


26.9 


30 


6 


8 


54 


7-49 


•925 


.646 


13.000 


18.27 


8.98 


.632 




623 


26.3 


30 





8 


71 


7.64 


.908 


.635 


14.000 


17- 59 


8.65 


.608 




600 


25-8 


29 


4 


8 


88 


7.80 


.891 


.624 


15.0 


16.93 


8.3. 


•585 




576 


253 


28 


8 


9 


06 7 . 96 


•875 


.613 



Owing to the increase of piston displacement per indicated 
horse-power, as shown in columns 8 and 9 of the table, some 
builders make the air cyhnders of compressors for mountain 
work of larger diameter for the same size of steam cylinder than 
those for sea-level service. As against the losses of the air end of 
the compressor at high altitudes, there is some gain in mean- 
effective pressure of the steam cylinders, because the exhaust 



* For this reason, in compressor-builder's catalogues, no account is taken of 
the ga^e pressures in tables cf compressor capacities at altitudes. 



AIR COMPRESSION AT ALTITUDES ABOVE SEA-LEVEL 195 

takes place against a lower atmospheric pressure. The same is 
true in part of the exhaust of machines using the compressed 
.air. But the resultant of these gains is small and cannot be 
given much weight in offsetting the losses. A large deduction, 
for example, would have to be made for the lower calorific power 
of a given fuel at high altitudes. 

The relation between compressor output and barometric 
pressure may be expressed simply in another way. Take the case 
of two compressors of the same size, one operating under an 
atmospheric pressure of, say, 14 lbs. and the other at 10 lbs. (cor- 
responding approximately to an altitude of 10,000 ft.) If the 
first compressor is producing 6 compressions, the final absolute 
pressure will be 14X6 = 84 lbs. or about 70 lbs. gage pressure. 
To produce the same gage pressure, the other compressor must 
work to an absolute pressure of 70+10 = 80 lbs., the number of 
compressions corresponding to which is f^ = 8. From each cubic 
foot of free air the first compressor will produce i of a cu.ft. of 
compressed air, and the second compressor, | cu.ft. Hence, 
the ratio of the respective outputs of the two compressors will 
be i-^| = f or 0.750. As compared with this, the ratio of the 
respective barometric pressures is 11 = 0.714. 

Mechanically Controlled Inlet Valves for High Altitudes. It is 
often stated that compressors the inlet valves of which are under 
mechanical control are of special advantage for work at altitudes 
above sea-level. While there is a measure of truth in this, the 
possible saving is necessarily small, except at considerable eleva- 
tions. The question presents itself as follows: If the valve 
resistance be diminished by introducing mechanical control, so 
that under normal conditions at sea-level the inlet air will begin 
to enter the cylinder a little earlier in the stroke, the volumetric 
capacity of the compressor is thereby increased. The loss of 
capacity due to resistance of the valve springs, etc., which has 
been assumed to be 0.75 lb. for ordinary poppet valves, is a 
constant, and therefore becomes proportionately of greater and 
greater consequence as the altitude increases, because its ratio to 
the diminishing atmospheric pressure goes on increasing. The 
percentage of saving obtained by eliminating the spring resist- 



196 COMPRESSED AIR PLANT 

ance, though small at or near sea-level, therefore becomes a 
matter of importance at great elevations; and the inlet valve 
which presents the smallest resistance to the entrance of the air 
into the cylinder will be the most economical for service in high 
mountain regions. 

Stage Compression at High Altitudes. According to the 
statement already made, the greater the altitude above sea-level 
the greater is the difference between the delivery pressure and 
atmospheric pressure; that is, the ratio of compression is greater. 
In Chap. \ the effect of clearance in the air cylinder was dis- 
cussed, and it is e\'ident that the percentage loss from this cause 
increases with the altitude, because the piston must advance 
farther before the clearance air has been re-expanded to a pressure 
below the diminished atmospheric pressure. Even if it be 
questioned whether it is worth while at sea-level to adopt stage 
compression for the ordinary pressures used in mining and 
tunnelling, the case is materially altered at high altitudes. 
For example, if it be desired to produce a gage pressure of 
75 lbs. at 5,000 ft. elevation, corresponding to an atmospheric 
pressure of about 12.2 lbs., 7.15 compressions are necessary. 
At sea-level this number of compressions would give a gage pres- 
sure of (14.7 X7.i5)-i4.7 = 90.4 lbs. So far as losses due to 
piston clearance are concerned, therefore, it is as reasonable 
to employ stage-compression for 75 lbs., at 5,000 ft. elevation, 
as for 90 lbs. at sea-level. In a compound compressor, too, 
it must be remembered that there is practically but one clearance 
space: that in the intake cylinder. The value of the intercooler 
also increases with the altitude, because, in beginning compression 
at an initial pressure below the normal, the greater total range 
of pressure through which the air must be carried involves the 
production of more heat. This additional heat must be effec- 
tually dealt with by the coohng arrangements, if loss from this 
cause is to be avoided. 

Considered from both the economic and thermodynamic 
standpoints, there can be no question as to the value of stage 
compression for high altitudes. There is not only a decrease in 
output and an increase in the cost of production of the air, due 



AIR COMPRESSION AT ALTITUDES ABOVE SEA-LEVEL 197 

to the added power required; but, as a result of these conditions, 
the compressor itself must be larger for a given output, and 
therefore its first cost will be greater than that of a compressor 
of the same capacity, working under normal atmospheric pressure. 
Hence, by introducing stage compression, a larger percehtage 
of saving is possible at high altitudes than at sea-leveL 



CHAPTER XIV 
EXPLOSIONS IN COMPRESSORS AND RECEI\TRS * 

Explosions in air compressors and receivers occur with suf- 
ficient frequency to demand careful attention. Though they 
are unquestionably attributable to ignition of volatile con- 
stituents of the lubricating oil, the immediate causes leading 
to this combustion are not altogether clear. But, since explosions 
occur only in dry compressors, some light may be thrown upon 
the subject by considering the conditions affecting the use of 
the lubricant. In Chap. V attention was called to the fact that, 
if the cylinder temperature of a dry compressor rises too high, 
not only does proper lubrication become difficult, but the oil 
itself may be decomposed. Ignition unattended by actual 
explosion is probably frequent; the discharge pipe near the 
compressor sometimes becomes red-hot, and ignition has even 
extended into the receiver without producing a destructive 
explosion. The discharge- valve chests and passages, and the 
pipe leading from the compressor to receiver, often contain a 
black, sooty residue from decomposition of the lubricant. But, 
on passing with the compressed air into the receiver, the volatile 
constituents of the oil thus Uberated would make a mixture of 
air and gas capable of producing an explosion. The extreme 
violence of such explosions is probably due in part to the high air 
pressure in the valve passages, discharge pipe, and receiver, 
since in high pressure air combustion is more active than in 
air at atmospheric pressure. 

As a number of the recorded compressor explosions have 
occurred at coUieries, the possible effects of the presence of 

* In connection with the revision of this chapter I have received valuable 
criticisms and suggestions from my friend Mr. C. M. Spalding, Mechanical Engi- 
neer with the General Electric Co. This help I desire gratefully to acknowledge 

198 



EXPLOSIONS IN COMPRESSORS AND RECEIVERS 199 

coal dust in the intake air of the compressor have been con- 
sidered. Such a deposit in the valve passages, together with 
the sooty residue from decomposition of the oil, might produce 
a condition favorable to an explosion. A spark caused by the 
friction of the compressor piston, if working dry, or, the con- 
tinual passage of air at a high temperature over the carbonaceous 
deposit, might produce spontaneous combustion, and ignite 
the inflammable mixture of oil-vapor and air.* However, there 
are enough cases where explosions have occurred at mines and 
works other than colleries to prove that explosions are not 
necessarily dependent upon the presence of coal dust in the 
intake air. When the compressor is improperly situated in a 
room close to the boilers, some coal dust might be present in 
the air; but, though possibly assisting in the explosion, the 
quantity could hardly be large enough to produce by itself the 
observed results. 

The primary cause of compressor explosions is undoubtedly 
to be found in the working conditions prevailing in the cylinder. 
In single-stage dry compressors very high temperatures are 
often reached, due to poor design of the air cylinder, or running 
too fast (as when the compressor is too small for its work), or 
attempting to produce too high a pressure. The temperature 
of the discharge air from a single-stage compressor is found by 
the formula given in Chap. X, 




in which: T and P are the absolute initial temperature and 
pressure of the intake air; T' and P^, the absolute final tem- 
perature and pressure; and n, the constant 1.41. Under 
normal conditions near sea-level, when the temperature of the 
atmosphere is 70° F., P = i4 lbs., and the gage pressure at dis- 
charge, 80 lbs., the final temperature would be: 

(O Q I -J- - \ 0. 29 
^ ^ ) =917° F. absolute, or 458° F. by 

the thermometer. 

* T. G. Lees Trans, Federated Inst. Mining Engineers, Vol. XIV, p. 568. 



200 COMPRESSED AIR PL\XT 

In using this formula, the compression is supposed to be 
purely adiabatic, no a4:count being taken of loss of heat by 
radiation or of any coohng effect from the water-jackets. Little 
heat can in any case be abstracted by the jackets of a single- 
stage compressor. Air is a poor conductor, and the volume 
in the cylinder is not long enough under the influence of the 
jackets to be much affected by them. In compressors of this 
type the chief office of the jackets is to keep down the tem- 
perature of the cyhnder walls and prevent the lubricating oil 
from being carbonized. It is probable that in a single-stage 
dry compressor, even if well designed and in good order, the 
discharge temperature generally mnges from 375° to 425° F., 
and may go higher. 

In \'iew of these considerations the quality of the lubricating 
oil used in the air cylinder, and especially its flashing- and 
ignition-points, are matters of hnportance.* The flashing-point 
of ordinary cyhnder oil may be taken as from 33o°-425° F. 
" An average of determinations on 40 samples of hea^y oils 
ha\ing an average flash-point of 360° F., gave an average burning- 
point of 398° F. High flash-test cylinder oils, from 50o°-56o° F., 
gave burning-points of 6oo°-63o° F."'t Common lubricating 
oils flash at about 250° F., and kerosene, sometimes carelessly 
used for cleaning valves, at 150^ F. or below. In the case of 
one explosion the flash-point of the cyhnder oil was found to 
be only 295° F.i It would appear, therefore, that an explosion 
in a compressor cyhnder, directly traceable to decomposition 
of the lubricant, is possible under normal conditions only when 
inferior, hght mineral oils are used. 

To produce an explosion there must be a sufficient increase 
of temperature to cause ignition of the lubricating oil or other 
combustible. In endeavoring to account for abnormal com- 
pressor temperatures, dift'erent theories have been advanced. 

* The flashing-point of oil is the lowest temperature at which it gives ofif com- 
bustible vapx)rs in sufficient quantity to be ignited by contact with flame. The 
ignition-point is the temperature to which the vapors must be raised in order to 
continue to bum. 

t Alex. M. Gow, Engineering Nrics, March 2d, 1905, p. 221. 

X John Morison, Trans. North of England Inst. Min. Engs., Vol. XXXVIII, p. 6. 



EXPLOSIONS IN COMPRESSORS AND RECEIVERS 201 

Some engineers have held that high cyHnder temperatures 
may result from leakage of delivery valves, or past the piston; 
the argument being that the hot, high-pressure leakage air 
raises the initial temperature of the cylinderful of air to be 
compressed on the next stroke, so that the final temperature 
becomes abnormally high. It would appear that this reasoning 
does not take account of the fall in temperature due to re-expan- 
sion of the leakage and clearance air behind the piston on the 
intake stroke. As it may fairly be assumed that the compression 
cycle is approximately adiabatic, the fall in temperature of the 
re-expanded air (disregarding the heating effect of the hot 
cylinder surfaces) would nearly correspond to the original rise 
of temperature due to the compression of this air. This theory, 
therefore, does not seem tenable. Other causes may possibly 
exist, but we have no definite knowledge, in fact, as to what 
does take place in an air cylinder working hot enough to pro- 
duce an explosion. In the absence of exact data, some light 
on the subject may be obtained from a study of the following 
examples : 

Examples of Explosions. An explosion which took place 
in one of the receivers of a compressor at the Chfton Colliery, 
England, attracted much attention, and is so instructive that 
some of the details are given here. * The air from the compressor 
passed to a series of 3 large receivers, the first being 7 ft. diameter 
by 40 ft. long. While running apparently under normal con- 
ditions the safety valves of the receivers suddenly began blowing 
off with a deafening roar. Flames several feet high issued at 
great pressure from the safety valves, and sparks were blown 
out at the joints of the 8-in. pipe leading from the compressor 
to the first receiver. The air main near this receiver was nearly 
red-hot. That the receivers did not burst was thought to be 
due to the relief afforded by the 4 safety valves — 2 on the first 
receiver and i on each of the others — and to the fact that the 
underground engines driven by compressed air continued run- 
ning for some minutes after the compressor was stopped. On 
examining the first receiver, after it had cooled, it was found 

* T. G. Lees, Trans. Federated Inst. Mining Engineers, Vol. XIV, pp. 555-559. 



202 cx):mpressed air plant 

that, just below the point where the air entered from the com- 
pressor, a mass of black carbonaceous matter had been deposited, 
from I J to 2 ins. thick and 6 sq. ft. in area. On analysis this 
showed: volatile matter, 55.8^, fixed carbon, 37.3%, and ash, 
6.9%. The material was charred and had the appearance of 
hard \'idcanite. A thin coating was noticed on the sides of the 
receiver (though only near the inlet pipe) and also in the pipe 
itself. The other two receivers were free from deposit. A 
carbonaceous coating, to a thickness of J in. was found on the 
discharge valves and passages. The cyhnder and piston sur- 
faces were not dry and, though they showed signs of excessive 
heat, were uninjured. 

The gage pressure was usually 60 lbs., which, vdih adia- 
batic compression, corresponds theoretically to a final tem- 
perature of 405° F., the temperature of the intake air from the 
engine-house being 80°. The lubricating oil used was guaranteed 
to have a fiash-point of 554". and ignition-point of 600° F. 
As the cylinders were water-jacketed, the discharge air should 
not, in regular working, reach these temperatures; in fact, 
readings pre\'iously taken from a thermometer in the outlet 
pipe showed that it usually registered about 350° F. It is 
significant, however, that on a pre\'iou5 occasion the mercury 
rose above 500°, and as the thermometer tube burst, the tem- 
perature at the time of the explosion was not known. After- 
ward a p)Tometer was fixed on the outlet pipe close to the 
discharge valves, and the temperature was found to rangf' 
generally from 400^-420° F., var}'ing ^viih the speed of the 
engine and the air pressure produced. Even with these tem- 
peratures, high as they are, it would seem impossible that 
ignition of the lubricating oil used could take place. It is evident 
that an unusual increase of temperature in the air cylinders 
must be accounted for, but no satisfactory explanation of this 
explosion has been offered. 

A \'iolent explosion occurred in the discharge pipe of a 
4-stage Laidlaw-Dunn- Gordon compressor, at a plant of the 
H. C. Frick Coke Co., Browmfield, Pa., which was furnishing 
air at 1,000 lbs. pressure for air locomotives. The compressor 



, EXPLOSIONS IN COMPRESSORS AND RECEIVERS 203 

/ 

was not damaged, though a large hole was blown in the pipe. 
It was thought that too much cylinder oil had been used, the 
record showing the consumption during the 5 months preceding 
the explosion to be 12 gals, per month. The average for the 
preceding year was 52.2 gals., but the reduction, great as if was, 
seemed to have been insufficient.* 

Evidence as to another cause of trouble was obtained when 
a second explosion in the same compressor took place two 
years later. A recording thermometer, which had been in- 
stalled in the discharge pipe close to the compressor, generally 
registered from 230°-25o° F., seldom exceeding 240°. A fusible 
plug, designed to blow out at between 325° and 350° F., was also 
set in the discharge pipe near the compressor. The monthly 
consumption of oil was further reduced to only 3.72 gals., a 
solution of castile soap and water being used almost exclusively 
for internal lubrication, with very good results. 

Previous to the second explosion, the compressor had been 
running normally. The day before, the maximum temperature 
was 240° F., the thermometer generally registering between 
190° and 230°, On the day of the explosion, the temperature 
reached 250° between 8 and 9 a.m. By ii a.m. it was evident 
that something was wrong, the temperature almost reaching 
270° at 1 1. 1 5. Investigation showed that the fourth-stage 
discharge valves were out of order, but the engineer thought 
that by careful running he could finish the day. He held the 
temperature between 250° and 265° until 2.50 p.m., when the 
explosion occurred; the chart of the recording thermometer 
then showing 270°, followed by a high peak in the curve. Coin- 
cident with the explosion, the fusible plug melted and blew out, 
releasing the pressure and checking the temperature at 620°. 
The compressor, which was uninjured, was stopped, and a new 
plug put in, taking about 15 mins., during which time the tem- 
perature dropped to 245°. On starting again (in doing which 
the engineer assumed an unnecessary risk) the temperature 
rose to 270°, before the compressor was shut down at 4.10 p.m. 

* It is probable that gummed oil and carbonaceous deposit had accumulated 
liberally wherever it could lodge in the interior of the compressor. — R. P. 



204 COMPRESSED AIR PLANT 

New valves and seats were put in, and on starting again 2 days 
later the temperature ranged from 2 2o°-24o°.* 

This explosion was ascribed by the management to " churn- 
ing " of the air, due to leaky discharge valves, which allowed the 
high-pressure air to re-enter the fourth-stage cylinder. It is 
possible, however, that the valves of the cyHnder were not 
working at all when the explosion took place, in which case the 
compressor became temporarily a 3-stage machine. If this be 
true, in compressing to 1,000 lbs. in three cylinders, the com- 
pressor was working under conditions for which it was not 
designed, and for which the cooling arrangements of the three 
remaining cylinders were inadequate. This explanation does not 
appear to be unreasonable. 

During the construction of the Xew York Aqueduct a fire 
occurred in a compressor receiver at one of the shafts. The air 
pressure was 80-90 lbs., and the receiver, set outside of the engine- 
house, was exposed to the hot sun. Part of the discharge pipe 
leading to the receiver became red-hot. On stopping the 
compressor and coohng down the receiver, the entire inner 
surface of the latter was found to be coated with carbonaceous 
matter at least J in. thick. Further investigation brought out 
the fact that the poppet discharge valves had sometimes 
occasioned trouble by sticking, and the engineer had been in the 
habit of using a squirt-can of kerosene to cut the gummy material 
clogging them. As the kerosene had a low flash-point, it was 
quickly vaporized, and when the cylinder temperature reached 
a sufficiently high point the explosion took place. 

In a case at Butte, Mont., two duplex compressors, with 
air cyhnders respectively of ^2^ by 60 ins. and 24 J by 48 ins., 
and running at 50 revs, per min., were forcing air at 80 lbs. 
pressure through a single 8-in. pipe. As somewhat over 1,200 
cu.ft. of compressed air per min. were being produced, the 
velocity of flow would be nearly 3,500 ft. per min., or 58 ft. per 
sec. It had been noticed several times that a portion of the 
discharge pipe close to the compressor became red-hot. In 

* Abstracted from a paper in Mines and Minerals, Vol. XXXII, p. 651, by 
William L. Affelder, Gen. Mgr. Bulger Block Coal Co., Bulger, Pa. 



Explosions in compressors and receivers 205 

the pipe between the compressors and receivers were several 
sharp bends, which increased the friction due to the rapid 
flow of the air. The receivers were always extremely hot. 
On one occasion the shaft timbering, 40 or 50 ft. below the 
shaft mouth, took fire from the hot air pipe. 

Although the observed results of this explosion were locaHzed 
in the discharge pipe, it is probable that oil was first vaporized 
either in the cylinder or when the compressed air was passing 
through the delivery valves; that a portion of it became hot 
enough to ignite, and in turn ignited an accumulation of oil 
vapor in the discharge pipe to the receiver. 

It seems necessary to hold that the primary cause of explosion 
is to be looked for in the cylinder, not in the discharge pipe or 
receiver. That is, it is reasonable to assume that the con- 
ditions leading to explosion are initiated at the point of maximum 
pressure (and therefore of maximum temperature), which is 
towards the end of the stroke and while the air is passing through 
the discharge valves. If this temperature is high enough, 
vaporization of some of the lubricating oil will occur, followed 
by ignition, which might extend into the mixture of air and oil 
vapor in the discharge pipe. 

Foul or poisonous gases may result from ignition of the 
lubricant in compressors or receivers, not necessarily followed by 
actual explosion. In an article in the Trans. Amer. Inst. Min. 
Engs. ,Yo\. XXXIV, p. 158, an instance is noted of combustion 
in an air pipe and receiver. The compressed air was being used 
in an imperfectly ventilated upraise in a mine, 1,200 ft. from 
the compressor, and two men lost their lives, while four others 
barely escaped asphyxiation. 

Other more or less similar cases are familiar to most miners, 
where foul air from the exhaust of machine drills has been 
observed; sometimes merely disagreeable, though often actively 
deleterious. The use of poor cylinder oil is probably responsible 
for this, as its lighter constituents may begin to volatilize and 
burn at a normal working temperature. Even if not actually 
fried on the hot metal surfaces, a low-grade oil will undergo 
a slow combustion or oxidation, which may produce enough 



206 COMPRESSED AlR ^LAnT 

carbon monoxide to raise materially the percentage of that 
poisonous gas in the confined atmosphere of working places 
of mines. 

Mode of Using Lubricant for Air Cylinders of compressors. 
Sight-feed lubricators, as commonly employed for steam cylin- 
ders, are best. On the Clifton Colliery compressor, mentioned 
above, ordinary oil-cups were used, holding about | pint; they 
were filled 4 times per day of 10 hours. With these oil-cups, if 
improperly adjusted, it would be possible for all the oil to be 
sucked into the cyHnder within a few strokes after being filled. 
Such a result might be inferred, indeed, in this case, because of 
the large quantity of carbonaceous matter — oil, coal dust, etc. — 
found in and around the discharge valves and in the receiver. 
The oil feed should be carefully regulated, and a smaller quantity 
used in an air cylinder than a steam cylinder of the same size — 
say, one-third as much. Excess of oil increases the tendency 
to gum the valves. For stage compressors of ordinary size, 
I drop of good cylinder oil every 4-5 minutes is sufficient. 

The periodical use of soap and water (soap-suds) is recom- 
mended for any compressor that cannot be shut down at short 
intervals for overhauling. It is fed into the air cylinder through 
an oil-cup, say during one day per week. Or it may be forced 
in by an oil-pump, with which the compressor should be pro- 
vided. Soap and water is a poor lubricant, and must be used 
more freely than oil, but it is effectual in cleansing the cylinder, 
valves, and ports from carbonaceous or gummy matter. If 
the compressor is to be stopped, as at the end of a shift, the 
feeding of soap and water should be discontinued some time 
before shutting down, and the oil-feed resumed, to avoid forma- 
tion of rust. Every compressor should be overhauled from 
time to time, and thoroughly cleaned in all parts, especially 
around the valves and passages, capable of furnishing a lodg- 
ment for oil or partly oxidized carbonaceous material. 

Precautions for Preventing Explosions: (i) Always inclose 
the inlet valves in a cold-air box, connecting with the outside 
air, to avoid taking air from the hot engine-room. This con- 
duces to economy in working, and by keeping down the final 



EXPLOSIONS IN COMPRESSORS AND RECEIVERS 207 

temperature tends to prevent decomposition of the oil. (2) The 
largest possible area of cylinder surface should be water- jacketed, 
including the cylinder heads. A liberal supply of the coldest 
water obtainable should be used for the jackets. The advantages 
in this respect of employing stage compression, with large inter- 
and aftercoolers, are undoubted. (3) Use only the best cyHnder 
oil, with high flash- and ignition-points and in as small quantity 
as is consistent with proper lubrication. (4) Keep the valves 
clean. In the design of the compressor there should be no 
recesses or pockets, around the valves or passages, where oil 
could accumulate. (5) Never introduce kerosene into the 
cylinder for cleaning the valves while the compressor is running. 
(6) Arrange the air intake so that coal dust will not be drawn 
into the cylinder with the inlet air. (7) Place a thermometer 
in the discharge pipe, close to the cylinder, so that the engineer 
can watch the temperature, and stop or slow down the com- 
pressor if the temperature of the discharge air rises too high. 
A continuously-recording thermometer is to be recommended. 

Conclusions. Though compressor explosions are not uncom- 
mon, it is undoubtedly true that, while mixtures in certain 
proportions of air and oil vapor are explosive, oil is often burnt 
in the cyhnders without causing a destructive explosion. This 
is proved by the frequent presence in air cylinders of sooty, 
carbonaceous deposits. The theories aiming to account for the 
observed phenomena by placing the responsibility entirely 
on leakage of discharge valves, or on *' churning " of the air 
back and forth in the cylinder, due to sticking of the valves, 
are not conclusive nor satisfactory for the reasons stated on p. 201 
regarding the adiabatic compression cycle. The whole subject 
is at present obscure. 

Before attempting to formulate conclusions, it would be 
desirable to secure more data. Definite knowledge of the 
conditions leading to explosions can be obtained by making 
laboratory investigations under controlled conditions, and 
applied to the circumstances which have been assumed to cause 
explosions. In this way, it could be determined to what extent, 
if any, the cylinder temperature is raised by leaky discharge 



208 COMPRESSED AIR PLANT 

valves, or by " churning " of the air. The effect of small 
particles of relatively non-conducting material, as carbon 
(coal dust) or lint, drawn into the cylinder with the intake air, 
could also be investigated. Such carbonaceous points might 
become incandescent, due to heating from small jets of flame 
from burning oil, and remain so long enough to ignite larger 
volumes of mixed air and oil vapor. It may be suggested that 
research leading to fuller knowledge of this subject might be 
carried out in the mechanical engineering laboratories of a 
university, or at the works of some compressor builder. 



CHAPTER XV 

AIR COMPRESSION BY THE DIRECT ACTION OF FALLING 

WATER 

Principle. When air in small bubbles is intimately mixed 
with water, the water breaks into foam, through which the 
bubbles tend to rise and escape. - But if the mixed air and water 
be drawn downward by a strong falling current, as in a vertical 
pipe, the air is compressed. Then if, after reaching the depth 
and head of water-column necessary to produce the compression 
desired, the direction of flow be changed to the horizontal and 
the velocity diminished, the bubbles of compressed air will be 
liberated and may be collected in a suitable chamber. The air 
pressure in this chamber corresponds to the effective head of 
water, that is, its depth below the level at the outflow or tail- 
.race. Thus, in this method of compression, no piston, valves, 
nor other moving parts, are used. 

As the bubbles are minute and thoroughly disseminated 
through the water during its descent, the total cooling surface 
is very large and isothermal compression results. The moisture- 
carrying capacity of the air is therefore smaller than if it were 
compressed adiabatically. During compression the percentage 
of moisture in each globule of air increases until the point of 
saturation is reached ; on further compression, the excess moist- 
ure is deposited, so that when re-expanded the air is relatively 
dry. This method was first tested on a working scale about 
1878, by J. P. FrizeU, of Boston, Mass.* 

Magog Plant. In 1896, the Taylor HydrauHc Air Com- 
pressing Co., of Montreal, erected a plant for the Dominion 

* Proceedings, Institution of Civil Engineers, London, Vol. LXTII, p. 347. 

209 



210 COMPRESSED AIR PLANT 

Cotton Mills, Magog, Province of Quebec* In a 128-ft. shaft 
(Fig. 114) was erected a vertical compressing pipe a, 3 ft. 8| in. 
diameter, the lower part increasing to 4 ft. 8 in., and made of 
i^-in. steel-plate. This pipe passes through the bottom of a re- 
ceiving chamber b, 12 ft. diameter by 12 ft. high, to which water 
is conducted from a head-race. Water flows into and fills the 
pipe, which extends nearly to the bottom of the shaft. Through 
a series of small feed pipes, air is drawn with the water into 
the top of the main pipe and is compressed while being carried 
down the shaft. The compressed air collects in a chamber c, 
while the water is returned to a tailrace near the top. The 
difference of level between intake and tailrace is about 22 ft., 
which produces the requisite speed of flow. Into the top of 
the vertical pipe a is inserted a telescoping section of pipe d 
(Fig. 115), carrying a bell-mouth e and headpiece /, terminating 
below in an inverted conoid g. Between e and g is an annular 
opening, through which the water enters the compressing pipe. 
The headpiece carries thirty 2-in. pipes h,h, 4 ft. long, open at the 
top and closed at the bottom. Into each of these pipes are 
screwed 32 short horizontal f-in. pipes i,i, all directed into the 
annular opening between e and g. As the entering water passes 
among the small pipes air is entrained, carried down the main 
pipe in the form of bubbles, and is thus compressed. 

Near the bottom of the shaft the compressing pipe enters 
the "separating" chamber c, 17 ft. diameter and 12 ft. high, 
open below and supported upon legs which raise it 16 ins. above 
the shaft bottom. Within this chamber is a conoidal " dis- 
perser " j, 12 ft. diameter. Below is an apron /, 5 ft. wide. 
When the water, charged with air bubbles, reaches the disperser 
it is first directed outwards, then deflected by the apron toward 
the center, and finally escapes through the open bottom of the 
separating tank into the return column. During this process 
of travel the compressed air separates from the water, most 

* The folio .ving description is based on an article in the Canadian Engineer, 
March, 1897, and information furnished to the author by the builders. See also 
Eng. and Mining Jour., Dec. 26th, 1896, p. 606, and Railway and Engineering 
Preview f Sept. 17th, 1898, p. 513. 



AIR COMPRESSION BY ACTION OF FALLING WATER 211 



=^ 




Fig. 114. — Taylor Hydraulic Air Compressor. 



212 



COMPRESSED AIR PLANT 



of it collecting in the upper part of chamber c. Part of the air 
is not Hberated at once, but collects in the annular space under 
the apron, and joins the main body of air through the pipe m. 
The pressure in the air chamber is due to the height of the re- 



Plan of Spider 

of Cylindrical 

Head Piece 




Fig. 115. 



turn water column in the shaft. The air is drawn off through 
the air main, alongside of the water column a. As the air bub- 
bles are surrounded by cold water, perfect isothermal compression 
is attained, with its corresponding advantages in minimizing the 
amount of moisture carried off in the air. 



^^ 




Fig. ii6.— Hydraulic Air-Compressing Plant at Kootenay. 

To face page 213. 



AIR COMPRESSION BY ACTION OF FALLING WATER 213 





Table 


XIV. 


Tests on the Magog Plant. '^ 




No. of 
Test. 


Water 

Discharged, 

Cu.ft. per 

Min. 


Available 

Head, 

Ft. 


Available 
Horse- 
Power. 


Air 

Delivered, 

Cu.ft. per 

Min. at 

Atmos. 

Pressure. 


Air 

Pressure, 

Lbs. per 

Sq. in. 


Actual 
Horse- 
power of 

Com- 
pressor. 


Efficiency, 
Per Cent. 


I 


6l22 


21.4 


247.7 


1377 


52 


132.5 


53-5 


2 


S504 


21.9 


228.0 


1363 


52 


131. 


57-5 


3 


4005 


22.3 


168.9 


1095 


52 


1^5-3 


62.4 


4 


7662 


21. I 


305-9 


1616 


52 


155-4 


50.8 


5 


6312 


21.7 


260.0 


1506 


52 


144.8 


55-7 


6 


7494 


21 . 2 


299.8 


1560 


52 


150.2 


50.1 



Temperatures during tests: external air, 75-80°; water, 75.2-80°; 
compressed air, 75.2-80°. 

The parts were incorrectly proportioned in this first instal- 
lation, and the efhciency could be increased by using a larger 
air chamber, to prevent air from going to waste. 

The theory of this mode of compression is as follows: The 
combined specific gravity of the mixture of air and water in the 
compressing pipe is less than that of the water in the return 
column. Therefore, the head required to overcome friction 
and to produce flow must be greater than if the apparatus were 
merely an inverted siphon, and as the difference in weight 
increases with depth (and air pressure produced) the motive 
head, or difference in level between the surfaces of water at 
inlet and in tailrace, must be correspondingly increased. 

Kootenay Plant. In 1.898-1900 another plant was built 
for the Kootenay Air Supply Co., Ainsworth, B. C. The topo- 
graphical conditions are such that a high head is obtained with- 
out sinking a deep shaft. From a dam the water is carried 
in a wooden-stave pipe, 5 ft. diameter and 1,354 ft. long, over 
a short trestle, built against the side of a gorge, to the receiving 
tank. The latter, 17 ft. diameter by 20 ft. high, is placed on a 
wooden tower, no ft. high (Fig. 116). From the tank the pres- 
sure pipe, T,T, ins. diameter, descends to the ground level and 
then down a shaft 105 ft. deep.f After compressing the air 

* Tests made by Prof. C. H. McLeod, of McGill University, August, 1896. 
Published in Eng. and Min. Journal, December 26, i8g6, p. 606. 
^Canadian Electrical News, September, 1898, p. 176. 



214 



COMPRESSED AIR PLAXT 



the water returns up the shaft to the taihace at the creek level. 
Fig. 117 shows the details of the recei\ing chamber at the 
bottom of the shaft. 



^^^^s» 




END ELEVATION 




SIDE ELEVATION 




PLAN 



Fig. 117. — Hydraul'c .\ir-Compressor at Kootenay. 



AIR COMPRESSION BY ACTION OF FALLING WATER 215 

The effective compressing head is 107 ft., the total height 
of the intake pipe being over 200 ft. This produces a high 
velocity of flow and large delivery of compressed air. The 
compressed air main, 9 ins. diameter, is 2 miles long, carrying 
from 4,200 to 4,600 cu.ft. of free air per min. Branch service 
pipes convey the air to neighboring mines, where it is used for 
rock-drills and other machinery. On the basis of 600 H.P., 
represented by the volume and pressure of the air, the cost 
of the entire plant, including pipe lines, was about $100 per 
horse-power. 

Victoria Plant was completed in 1906 at the Victoria Copper 
Mine, Rockland, Ontonagon Co., Mich. The local conditions 
led to a novel mode of installation. The water is conducted 
from a dam on the Ontonagon River through a 4,700-ft. canal, 
furnishing a head at the terminal forebay of 72 ft. above the 
river-level. Three independent units are built side by side at 
19-ft. centers in a vertical shaft 340 ft. deep. In the original 
design, the subdivision of the air, as admitted at the intake 
head (Fig. 118) was carried farther than in either of the plants 
described above, by inserting 1,800 f-in. horizontal feed pipes, 
in the series of larger vertical pipes encircling the inverted cone. 

After the plant was put in operation, serious trouble was 
experienced by the freezing up of the small pipes of the intake 
heads, due to the severe winter climate of the region. This 
led to the removal of the heads, the water being allowed simply 
to flow into the top of the compression pipes. The breaking up 
and agitation of the mass of water, in changing its direction of 
flow from the forebay into the compressing pipes, entrained the 
air quite efi&ciently, and it is stated that the capacity of the 
plant, in cubic feet of free air compressed per minute, is prac- 
tically the same as when the intake heads were in use. 

The compressing pipes are 5 ft. diameter, lined with con- 
crete, and separating cones and dispersers, also of iron and 
concrete, are built in a chamber at the bottom. In this chamber, 
281 ft. long and 18 ft. by 21 ft. average cross-section, the com- 
pressed air is trapped and thence drawn off through a 24-in. 
main. The compressing water, flowing down the intake pipes, 



216 COMPRESSED AIR PLANT 

stands normally at a level about 14 J ft. below the roof of the 
chamber, thus leaving an air capacity of about 80,000 cu.ft. 
Connected with the end of the air chamber is an inclined shaft, 
270 ft. in vertical depth, through which the water returns to 
the surface. The tailrace from this shaft is 72 ft. below the level 
of the intake, this height measuring the motive head producing 
the flow of water. Thus the air in the underground chamber is 
under a pressure due to 270 ft. head of water, or 118 lbs. sq. in. 

For regulating the operation of the original plant a pipe 
passed from the air chamber up the compressing shaft to the 
surface, whence branches were led to the intake heads. The 
compressed air conveyed in this regulating pipe operated a 
device connected with each intake head, whereby the latter was 
automatically raised above the water-level in the receiving tanks 
whenever the air pressure exceeded the normal, thus stopping 
the flow of air through the -feed pipes. A 12 -in. blow-off pipe 
passes from the water-level in the air chamber to the mouth 
of the inclined shaft carrying the return water column. If air to 
the full compressor capacity is drawn off, the water-level in the 
air chamber rises as the air pressure falls, thus sealing the lower 
end of the blow-off* pipe; then, when the consumption of air 
decreases, the pressure in the chamber rises, depressing the 
water-level until the blow-oft' orifice is uncovered, when more air 
is blown off. Thus the working pressure is maintained within 
quite narrow limits. The great size of the air chamber — cor- 
responding to the receiver of an ordinary air-compressor — gives 
a large storage capacity. 

When all 3 compressing units are in operation, with a total 
capacity of from 34,000 to 36,000 cu.ft. of free air per min., 
about 70,000 cu.ft. of air per min. may be drawn off for a period 
of 18 minutes, without causing a drop in pressure of more than 
5 lbs. For each unit, the output ranges from 9,000 to 12,000 
cu.ft. per min., and the volume of water used from 12,700 to 
14,800 cu.ft. Tests made on a single intake head in May, 1906, 
by Prof. F. W. Sperr, gave the following results:* 

* For further details see article by D. E. Woodbridge, Eng. &* Mijt. Jour., 
Jan. 19, 1907, p. 125. Also, A. II. Rose, Mines b° Min., Mch., 1907, p. 346. 



^^^^ 



3 of these onitB 
side by side, 
19 ft center to c« 
erected in one eJ 




Fig. ii8. — Hydraulic Air-Compi 



^^^^^ 



Air Pipe to Miue and Mill 
Pressure 118 lbs pen eq in" 



^^ 




^; 



Plant. Victoria Mine, Mich. 

To face page 216. 



b^^^x^$;^;;gj a^ai^yx<^^y/xc^>^.^.^<s^-^^^^ 



-86g 



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1^ 







I- 



LLl 

_J 




ho2-^ 



a. 








tX) 






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x: 


o 




(U 


hs 




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AIR COMPRESSION BY ACTION OF FALLING WATER 217 



Table XV. — Air Measurements 





Velocity, 
Ft. per Sec. 


Cu.ft. per 
Min. 


Absolute 


Pressures. 




Sq.ft. 


Free Air, 
Lbs. 


Compressed 
Air, Lbs. 


Horse-power. 


4 
4 
4 


44-09 
49-74 
38.50 


10,580 

11,930 

9,238 


14 
14 
14 


128 
128 
128 


1,43^ 
1,623 
1,248 



Water Measurements 



Flume 
Area. 


Velocity, 
Ft. per Sec. 


Cu.ft. per 
Min. 


Head, Ft. 


Horse-power. 


Efficiency, 
Per Cent. 


71-75 
67.03 
72. 16 


3 033 
3-684 
2.936 


13,057 
14,820 
12,710 


70.5 
70.0 
70.6 


1,741 
1,961 

1,700 


82.17 
82.27 
73-50 



The air is used at the Victoria Mine for general power pur- 
poses at the mine and mill, including a 500-H.P, hoisting-engine, 
and 7 pumps. The cost per horse-power is about $2.25 per 
year, including all operating expenses. Over 4,000 H.P. can be 
developed by the 3 compressing units. 

Cobalt Power Co.'s Plant. Following a series of efhciency 
tests made in 1909 on a large number of steam compressors 
at the silver mines in the Cobalt, Ont., district (see Chap. X, 
p. 152-168), a Taylor compressor was built at Ragged Chutes, 
9 miles from Cobalt, on the Montreal River. In a distance of 
1,000 ft. there is a drop of 54 ft. From the forebay the water flows 
into two i6-ft. heads (Fig. 119), in each of which sixteen 14-in. 
vertical intake pipes are set in a horizontal disk. Below the 
disk the heads taper to 8 ft. 4I in., below which point they 
extend 15 ft., telescoping into the tops of 8J-ft. diameter con- 
creted compression shafts, 330 ft. deep. To regulate the inflow, 
and adjust the position of the heads to the forebay water level, 
the heads are suspended from 2 vertical hydraulic cylinders. 
The heads, with their large diameter intake pipes, are designed 
to prevent freezing in the severe winter chmate of the region 
(see above, under Victoria Plant). 



218 COMPRESSED AIR PL.\XT 

The water with the entrained air flows through the heads 
at a velocity of 15-19 ft. per second. Due to the compression 
of the air in the shaft, this velocity gradually diminishes, with 
a further reduction in the lower 40 ft. of shaft, which is flared 
to 12J ft. diam. At the bottom of each shaft is a steel-plate 
capped concrete diverting cone see also Fig. iiS and accom- 
paming description < . 

The shafts terminate in a tunnel (air chamber), 26 ft. high. 
20 ft. wide and i.ooo ft. long, in which the air is completely 
hberated. and which series as a receiver. So great a length 
was not required for these purp>oses, but was adopted to utilize 
the total head (54 ft.) of the steam. From the tunnel, air is 
drawn off through a 24-in. pipe passing in an incHned riser to 
the vertical tail shaft. The 12-in. blowoff pipe acts in case 
the air pressure in the tunnel should force the water level below 
the roof of the outlet to the tail shaft, and so cause fluctuations 
of pressure. The air pressure produced is that due to the 
net head of water in the tail shaft; in this case, 276 ft., cor- 
responding to 1 20 lbs. per sq. in. 

This plant compresses 40.000 cu.ft. free air per min., cor- 
responding to about 5.500 HP. The air is carried through 9 
miles of 20-in. pij>e to Cobalt. From there branch pip)es con- 
nect with the different mines, the total piping (20, 12, 6 and 
3-in.) being about 2 1 miles. Total cost, excluding piping, about 
$1,000,000. or $185 per H.P. The air is sold by the company 
at 25 cents per i.ooo cu.ft. at 100 lbs. pressure.* 

Other Plants. In the State of Washington there is a 2C0-H.P. 
plant. Head of water. 45 ft., height of compressing pipe, 260 
ft., diameter, 3 ft.; volume of water. 53 cu. ft. per second; air 
pressure. 85 Ibs.t 

A small plant at Peterborough. Ont., has an i8-in. com- 
pressing pil>e, in a 42-in. shaft. Depth of separating chamber 
below discharge level, 64 ft., air pressure. 2>, lbs. 

Near Norwich. Conn., on the Shetucket River, there is a 
large plant for general power purposes. 1 

* C. H. Taylor. Mints ^ Min., Apl., 1910, p. 532. 
T Eng. ^ Min. Jour., Apl. 27, 1901. 
X Compressed Air Magazine, Apt, 1906, p. 3,980. 




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two 


o 


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d 

d 

2 
&? 

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fi4 



AIR COMPRESSION BY ACTION OF FALLING WATER 219 

In 1907-8 a hydraulic air compressor was installed at 
a silver mine at Clausthal, Germany. Fig. 120 shows the 
general design, with details of the intake head and compressing 
chamber. A flow of water in the tunnel t is led through an 
8|-in. cast-iron pipe a, to the air intake h, which consists of a 
number of flaring rings /, in the upper rim of each of which is a 
series of smafl holes k, for admitting the air. Additional inlet 
area is provided at the top of the intake by a nest of small 
curved pipes m. The mixed air and water pass into the 8|-in. 
compression pipe c, 492 ft. long, laid in an inclined shaft, and 
discharging into the compressing chamber d. This chamber 
is 52 in. by 14 ft. 9 in. high. From a point near its top the 
compressed air passes through pipe n to the automatic check- 
valve e, and thence, by pipe h, to the receiver i. The water 
leaves the compressing chamber by the 8J-in. pipe /, which 
discharges at a point 164 ft. above, into a tailrace occupying 
the mine level u. An equalizing discharge pipe g, from the com- 
pressing chamber, is led up the shaft, parallel to /, entering 
the latter at the level of the tailrace. Total cost of the plant 
is stated to be $3,750.* 

The average flow of water is 792 gals, per min. ; which, fall- 
ing through a vertical height of 325 ft., produces theoretically 
66.3 H.P. A flow of 845 gals, per min. gave 353 cu.ft. of air, 
at 71.2 lbs. gage. To compress i cu.ft. of air adiabatically 
to this pressure requires 0.147 H.P. and to compress 353 cu.ft., 
about 51.9 H.P, Since 70.5 theoretical H.P..are produced by 

the flow of 845 gals, per min., the efliciency is — '- =73.6%. 

70-5 

Modifications of the hydraulic compressor have been pro- 
posed: the McFarlane, described in Eng. and Min. Jour., Oct. 
10, 1908, and the Blakney, Eng. df Min. Jour., Apl. 24, 1909. 
(For a general description of hydraulic compressors, see Western 
Eng'g, March, 191 7.) 

The first cost of hydraulic air compressors is not excessive, 
while the maintenance and running expenses are very low, 

* Abstracted from a description by P. Bernstein, in Gliickuiif, March 14, 1908. 
Translation by E. K. Judd in Eng. & Min Jour., August i, 1908, p. 228. See 
also Zeitschr. ver. Deutsche^ Ing., Nov. 5, 1910. 



220 COMPRESSED AIR PLANT 

compared with those of ordinary compressors. No skilled 
attendance is required, and depreciation is nominal in substan- 
tially erected plants. By comparing the figures given in Tables 
XIV and XV, it will be seen that Victoria Plant gave a marked 
increase in efficiency, due to the greater motive head, and a 
more complete separation of the air from the water in the receiv- 
ing chambers. 

It has been suggested that it might be feasible to employ the 
system in connection with an ordinary compressor plant. That 
is, to produce a low air pressure by the water plant, and then to 
admit this air to the compressor cylinder where it would be 
brought to the required higher tension. In effect, this would 
be stage compression, in which the air would be cooled to normal 
temperature before entering the high-pressure cylinder. 

This system of air compression is generally unsuitable for 
small plants, as the first cost is large. 



Part Second 

TRANSMISSION AND USE OF 
COMPRESSED AIR 



CHAPTER XVI 
CONVEYANCE OF COMPRESSED AIR IN PIPES 

The diameter of the pipe is of vital importance, and when 
proportioned properly to the volume of air, and to the distance, 
the transmission losses are very small compared with the other 
losses incident upon air compression. Transmission losses 
appear in two ways: as loss of power, and as loss of pressure or 
head. 

Loss of Power. The large loss of power due to the heating 
of the air during compression and its subsequent cooling, has 
already been considered. This cooling takes place so quickly 
in the receiver and piping that the resulting loss is not properly 
chargeable to transmission. The air assumes the temperature 
of the surrounding atmosphere in the first few hundred feet, so 
that when conveyed to long distances the calculation for trans- 
mission loss may be made without regard to the effect of tem- 
perature upon the volume of the air. The power in the com- 
pressed air is due not only to its pressure, but also to its volume, in 
terms of cubic feet of tree air. While the pressure is reduced 
by frictional loss in transmission, this reduction is accompanied 
by a proportionate increase in volume, and a certain compensa- 
tion is produced. Although the pressure of the air at the motor 
is diminished, there is no loss in the final volume of free air. 
As shown below, the loss of pressure due to the conveyance 

221 



222 COMPRESSED AIR PLANT 

of air in pipes is small, but the actual loss of power is still smaller. 
The pipe itself acts in a measure like a receiver — as a reservoir 
of power. Much of the transmission power loss experienced 
in practice is due to leakage from joints and flaws in the pipe. 

Loss of Pressure for short distances takes place according 
to the laws governing the flow of all fluids, varying directly 
as the length of pipe, directly as the square of the velocity, and 
inversely as the pipe diameter. For long distances the applica- 
tion of these laws becomes somewhat complex. In addition 
to the factors just given, it is necessary to take into account 
the volume and pressure of the air, and the initial and final 
pressures at the ends of the pipe Hne. In general, for a given 
diameter of pipe, when the volume of air discharged and its 
initial pressure remain constant, the loss of pressure is pro- 
portionate to the length of the pipe. 

But in actual service the initial pressure and volume of 
discharge do not remain constant, and, in the passage of the air 
through the pipe, other modifying factors must be taken into 
account. In flowing through a long line of piping the pressure is 
gradually reduced by friction, while the volume is correspond- 
ingly increased. Therefore, to maintain in the pipe the flow 
of a given quantity of air, the volume of which is constantly 
increasing, the velocity also must increase, and this requires an 
increase of head or pressure. 

The formulas commonly used assume that the loss of head is 
proportional to the length of pipe, so that, if a certain head be 
required to maintain the flow of a given quantity of air in a 
pipe I, GOO feet long, twice this head would suffice for a pipe 
2,000 feet long. But, when the air has passed through the 
first thousand feet of pipe its motive head has been lost; and 
as the volume has thereby increased, a greater head will be 
necessary to maintain the flow in the second thousand feet. 
The ordinary formulas do not take into account the increase of 
volume due to this loss of head. To transmit a given volume 
of air at a uniform velocity and loss of pressure would require 
a pipe of gradually increasing area. This of course is impractic- 
able, and if the discharge is to be kept constant in pipe of uni- 



CONVEYANCE OF COMPRESSED AIR IN PIPES 223 

form section, both volume and velocity must increase as the 
pressure is reduced by friction. The loss of head in properly 
proportioned pipes is so small, however, that in practice the 
increase in volume is usually neglected. 

The discharge capacity of piping is not proportional to the 
cross-sectional area alone. Although the periphery is directly 
proportional to the diameter, the interior surface resistance is 
greater in a small than in a large pipe, because the ratio of 
perimeter to area is greater. To pass a given volume of air a 
I -in. pipe of given length requires over 3 times as much head as a 
2-in. pipe of the same length. The character of the pipe also, 
and the condition of its inner surface, have much to do with the 
frictional resistance. The irregularities incident upon coupling 
together the lengths of pipe also increase friction. As the 
influences by which the values of some of these factors may be 
modified are not fully understood, the results obtained from 
formulas are only approximately correct. 

D'Arcy's Formula, as adapted to compressed-air transmis- 
sion, is: 

^ j d'{pi-p2) ^ c\/^ I P1-P2 
D=c\^ , or D=—7=yX\ 

^ wil ' v I ^ wi \ 

in which: 

D = the volume of compressed air, cubic feet per minute, dis- 
charged at the final pressure; c = a coefficient varying with the 
pipe diameter, as determined by experiment; ^ = nominal diam- 
eter of pipe, inches;* / = length of pipe, ft.; pi and ^2 = initial 
and final gage presssures, lbs. per sq. in. ; wi = density of the 
air, or its weight in lbs. per cu.ft., at pressure p\. 

The formula in its factored form (see above) is convenient 
for use. Table XVI gives the values of c, d^, and c\/d^. 

Table XVII gives the value of wi for initial gage pressures up 

to 100 lbs. per sq. in.. Table XVIII the values of \— — -- for 
terminal pressures of 20-100 lbs., and pressure losses of i-io 

* The actual diameters of wrought-iron pipe are not the same for all sizes 
as the nominal diameters. This difference is small except in the ij-in. and i§-in. 
sizes, the actual diameters of which are 1.38 ins. and 1.61 ins. respectively. 



224 



COMPRESSED .\IRPL.\XT 



lbs.* Intermediate values are obtained by interpolation. Xo 
allowance is made for pipe leakage, nor for incidental friction due 
to bends in the pipe 'Table XXII;. 



Table XTVT 



Diartie:er :: ?:pe. 


Vaiues oi 


Fiith Powers 01 


Va-.-es :f 




.-.re;. 


c 


d 


\ > 




I 


45 5 


I 


Al ^ 




2 


52.6 


32 


297 




3 


56.5 


243 


876 




4 


58-0 


1.024 


1,856 




5 


59 


3.125 


3,298 




6 


59 8 


7,776 


5,273 




7 


60.3 


16,807 


7,817 




8 


60.7 


32,768 


10.9S8 




9 


61.0 


59.049 


14,812 




lO 


61.2 


100,000 


19,480 




II 


61.4 


161,051 


24.800 




12 


61.6 


248,832 


30,926 





Gage Pressure. 


^ 


fl 
x" ' 


Gage Pressure. 

L-s. 


ITl 


\ -X! 





0.0761 


0. 276 


55 


0.3607 


0.600 


5 


0.1020 


319 


60 


0.3866 


0.622 


10 


0.1278 


0.358 


6S 


0-4125 


0.642 


IS 


0-1537 


0.392 


70 


0.4383 


0.662 


ao 


0.1796 


0.424 


75 


0.4642 


6S1 


25 


0.2055 


453 


80 


0.4901 


0. 700 


30 


0-2313 


0.481 


85 


0.5160 


0.718 


35 


0.2572 


0-507 


90 


0.5418 


0.736 


40 


0. 2831 


0-532 


95 


0.5677 


753 


45 


0.3090 


0-556 


100 


0.5936 


0.770 


50 


3348 


0.578 \ 









Example: Given a 5-in. pipe. 2.500 ft. long; how many cu.ft. 
of air per min. at 70 lbs. initial pressure can be transmitted, with 
a loss of pressure of 3 lbs.? 



* Keproduced by permission from Compressed Air^ Feb., 1898, pp. 374-376, 



CONVEYANCE OF COMPRESSED AIR IN PIPES 225 



From Table XVI, cVd^ = s,2g8; from Table XVIII, yj^^—^ 

= 2.615 ^^^ ^ ^ =50- Substituting in the formula: 

_ 3,208 ^ - ... 

D= X 2.615 = 172.5 cu.ft. compressed air per mm. 

Volumes of compressed air may be converted into free air 
by multiplying by the absolute pressure in atmospheres (i 
atmos. = 14.7 lbs.). Thus, 100 cu.ft. of air at 80 lbs. gage, or 
94.7 absolute pressure, correspond to 644 cu.ft. of free air, at 
sea-level. Table XIII gives the air pressures in lbs. per sq.in. for 
altitudes to 15,000 ft., with the corresponding barometric 
readings. 

Graphic Solution of D'Arcy's Formula (C. W. Crispell, 
Trans. Am. Inst. Min. Engs., Vol. LVIII, p. 97.) The follow- 
ing problems may arise (see Fig. 121): 

1. To find the diameter of pipe; given the volume of com- 
pressed air, length of pipe, initial pressure and maximum drop 
in pressure. 

With a straight-edge, join the scales marked length of pipe 
and cu.ft. of compressed air, and note the intersection on axis A . 
Join the initial pressure with the drop in pressure, and note 
intersection on axis B. A line joining these two points of inter- 
section will cut scale No. 3 at the required pipe diameter. 

2. To find the volume of compressed air that a pipe will 
carry; given the length and diameter of pipe, .initial pressure, 
and maximum allowable drop in pressure. 

Join the initial pressure with the drop, and note the inter- 
section on axis B. Join this intersection with the pipe diameter, 
and note the intersection on axis ^ of a prolongation of this 
line. A line joining this point on A with the pipe length will 
cut scale No. 2 at the required volume. 

3. To find the maximum length of pipe that will carry a 
given volume of air; given the pipe diameter, initial pressure, 
and maximum drop. 

Join the initial pressure with the drop, and note intersection 
pn axis B. Join this point with the pipe diameter, and note 



226 



COMPRESSED AIR PLANT 



Table XVIII (William Cox) 

Values of -v ' - — — 



Final 








Losses 


5 OF Pressure, 


Pi -Pi. 










Pressure 
Pj. Lbs. 


I lb. 


2 lbs. 


3 lbs. 


4 lbs. 


5 lbs. 6 lbs. 


7 lbs. 


8 lbs. 


9 lbs. 


10 lbs. 


20 


2 325 


3-241 


3.918 


4.466 


4 930 


5-336 


5-693 


6.014 


6.309 


.— ^-^ 

6-574 


21 


2.293 


3.198 


3.868 


4 


.410 


4 


.870 


5.272 


5.627 


5 


■946 


6.237 


6.502 


22 


2. 262 


3-157 


3.819 


4 


-356 


4 


.812 


5-2II 


5-564 


5 


.878 


6.168 


6.432 


23 


2.233 


3-II7 


3-772 


4 


.304 


4 


.756 


5-152 


5.501 


5 


.814 


6. 102 


6.362 


24 


2.205 


3-079 


3-727 


4 


.254 


4 


.702 


5-093 


5 440 


5 


-752 


6.036 


6. 296 


25 


2.178 


3.042 


3.684 


4 


. 206 


4 


.649 


5-036 


5-381 


5 


.688 


5-973 


6.233 


26 


2.152 


3-007 


3.642 


4 


.15S 


4 


.597 


4.981 


5-323 


5 


-630 


5 913 


6.173 


27 


2. 127 


2.973 


3.601 


4 


. 112 


4 


-548 


4.928 


5.268 


5 


-572 


5-856 


6. 113 


28 


2.103 


2.939 


3-561 


4 


068 


4 


•499 


4.877 


5-215 


5 


.518 


5-799 


6 056 


29 


2.079 


2.907 


3-523 


4 


024 


4 


-452 


4.828 


5-164 


5 


.466 


5- 745 


5-C99 


30 


2.056 


2.876 


3-485 


3 


982 


4 


.408 


4.781 


5-114 


5 


.414 


5-691 


5-942 


31 


2.034 


2.844 


3-448 


3 


942 


4 


365 


4- 735' 5 066 


5 


•364 


5 637 


5.888 


32 


2.012 


2.815 


3-414 


3 


904 


4 


323 


4.690: 5.019 


5 


.312 


5-586 


5 834 


33 


1,991 


2.786 


3-381 


3 


866 


4 


282 


4.646 


4-971 


5 


. 264 


5-535 


5-782 


34 


I. 971 


2-759 


3 348 


3 


830 


4 


242 


4.603 


4.926 


5 


.216 


5.487 


5-733 


35 


I • 952 


2.733 


3-317 


3 


794 


4 


202 


4-561 


4.881 


5 


.170 


5.439 


5.686 


36 


1-933 


2.707 


3.286 


3 


758 


4 


164 


4.520 


4-839 


5 


. 126 


5-394 


5 639 


37 


1-915 


2.682 


3-255 


3 


724 


4 


126 


4.480 


4-797 


5 


.084 


5.349 


5-594 


38 


1.897 


2.656 


3-225 


3 


690 


4 


090 


4.441 


4-757 


5 


042 


5.307 


5 550 


39 


1.879 


2.632 


3.196 


3 


658 


4 


054 


4 404 


4-717 


5 


.002 


5.265 


5 • 509 


40 


1.862 


2.608 


3-168 


3 


626 


4 


02c 


4-368 


4.680 


4 


962 


5.226 


5.468 


41 


I 845 


2.585 


3-140 


3 


596 


3 


987 


4 333 


4-643 


4 


924 


5.187 


5-426 


42 


1.829 


2 . 563 


3 114 


3 


566 


3 


956 


4.299 


4.609 


4 888 


5.148 


5-385 


43 


1. 813 


2.542 


3-088 


3 


538 


3 


924 


4-267 


4-575 


4-852 


5.109 


5-344 


44 


1.798 


2.521 


3-064 


3 


510 


3 


895 


4-235 


4-540 


4.814 


5.070 


5-306 


45 


1.783 


2.501 


3-040 


3 


484 


3 


866 


4.203 


4-506 


4 778 


5 . 034 


5-268 


46 


1.769 


2.481 


3-017 


3 


458 


3 


837 


4-171 


4-471 


4-744 


4.998 


5-230 


47 


1-755 


2.462 


2-995 


3 


432 


3 


808 


4 139 


4-439 


4 


710 


4.962 


5-192 


48 


1.742 


2.444 


2.972 


3 


406 


3 


779 


4.109 


4.408 


4 


676 


4.926 


5.155 


49 


1.729 


2.426 


2.950 


3 


380 


3 


752 


4.080 


4-376 


4 


642 


4.890 


5.120 


50 


I. 716 


2.407 


2.927 


3* 


356 


3 


725 


4-051 


4-344 


4 


608 


4-857 


5-085 


51 


1.703 


2.389 


2 . 906 


3 


332 


3 


698 


4.022 


4-313 


4 


578 


4.824 


5-050 


52 


1 .690 


2.372 


2.886 


3 


308 


3 


671 


3.993 


4-283 


4 


546 


4-791 


5-015 


53 


1.678 


2-355 


2.865 


3 


284 


3 


645 


3.965 


4-254 


4- 


516 


4-758 


4.983 


54 


1.666 


2.338 


2.844 


3 


260 


3 


620 


3.938 


4.225 


4- 


484 


4.728 


4.952 


55 


1-654 


2.321 


2.823 


3. 


238 


3. 


596 


39" 


4.196 


4- 


456 


4.698 


4.920 


56 


1.642 


2.304 


2.804 


3- 


216 


3. 


571 


3.885 


4.169 


4- 


428 


4.668 


4.889 


57 


1.630 


2.289 


2.785 


3- 


194 


3- 


547 


3.860 


4-143 


4- 


400 


4.638 


4.860 


58 


1 .619 


2.273 


2.766 


3. 


172 


3. 


524 


3-835 


4. 117 


4- 


372 


4. 611 


4.832 


59 


1.608 


2.258 


2.747 


3- 


152 


3. 


502 


3-81T 


4-091 


4- 


346 


4.584 


4.80; 


60 


1-597 


2. 242 


2.730 


3 


132 


3. 


479 


3-787 


4.066 


4. 


320 


4.557I 


4-775 



CONVEYANCE OF COMPRESSED AIR IN PIPES 



227 



Table XVIII — Continued 

Values of 



V^-^ 



Final 








Losses of Pressure, 


pi —pi. 






Pressure 
p^, Lbs. 


I ib. 


2 lbs. 


3 lbs. 


4 lbs. 


5 lbs. 


6 lbs. 


7 lbs. 


8 lbs. 


9 lbs. 


10 lbs. 


6i 


1-586 


2. 228 


2. 712 


3. 112 


3.458 


3-764 


4.042 


4.294 


4.530 


4.747 


62 


1.576 


2. 214 


2.695 


3.092 


3-437 


3 


742 


4 


019 


4 


268 


4 


503 


4 


718 


63 


1.566 


2. 200 


2.678 


3.074 


3.417 


3 


720 


3 


995 


4 


244 


4 


476 


4 


693 


64 


1.556 


2.186 


2.662 


3.056 


3.397 


3 


698 


3 


971 


4 


220 


4 


452 


4 


.668 


65 


1.546 


2.173 


2.647 


3.038 


3.376 


3 


676 


3 


948 


4 


196 


4 


428 


4 


642 


66 


1-537 


2. 160 


2.631 


3.020 


3 356 


3 


654 


3 


926 


4 


172 


4 


404 


4 


617 


67 


1-528 


2.147 


2.615 


3.002 


2,-2>3>1 


3 


634 


3 


905 


4 


150 


4 


380 


4 


592 


68 


I-519 


2.134 


2. 600 


2.984 


3.318 


3 


615 


3 


884 


4 


128 


4 


356 


4 


566 


69 


1.510 


2. 122 


2.584 


2.968 


3.300 


3 


596 


3 


863 


4 


104 


4 


332 


4 


541 


70 


1. 501 


2. 100 


2.570 


2.952 


3-283 


3 


576 


3 


842 


4 


082 


4 


308 


4 


516 


71 


1.492 


2.098 


2-556 


2.936 


3.265 


3 


556 


3 


820 


4 


060 


4 


284 


4 


494 


72 


1.484 


2.086 


2-543 


2.920 


3.247 


3 


537 


3 


799 


4 


038 


4 


263 


4 


471 


73 


1.476 


2.075 


2-529 


2.904 


3.229 


3 


517 


3 


778 


4 


018 


4 


242 


4 


449 


74 


1.468 


2.064 


2-515 


2.888 


3. 211 


3 


498 


3 


759 


3 


998 


4 


221 


4 


427 


75 


1.460 


2.052 


2.501 


2.872 


3.193 


3 


480 


3 


741 


3 


978 


4 


200 


4 


405 


76 


1-452 


2.041 


2.487 


2.856 


3.177 


3 


463 


3 


723 


3 


958 


4 


179 


4 


383 


77 


1.444 


2.030 


2.473 


2.842 


3.162 


3 


446 


3 


704 


3 


938 


4 


158 


4 


361 


78 


1-436 


2.019 


2.461 


2.828 


3.146 


3 


429 


3 


686 


3 


918 


4 


137 


4 


339 


79 


1.428 


2.009 


2.449 


2.814 


3.130 


3 


412 


3 


667 


3 


898 


4 


116 


4 


317 


80 


1. 421 


1.999 


2.437 


2.800 


3. 115 


3 


395 


3 


648 


3 


878 


4 


095 


4 


294 


81 


1. 414 


1.989 


2-425 


2.786 


3.009 


3 


377 


3 


630 


3 


858 


4 


074 


4 


272 


82 


1.407 


1.979 


2.413 


2.772 


3.084 


3 


360 


3 


611 


3 


840 


4 


053 


4 


253 


83 • 


1.400 


1.969 


2.401 


2-758 


3.068 


3 


343 


3 


593 


3 


820 


4 


035 


4 


234 


84 


1-393 


1.959 


2.388 


2-744 


3.052 


3 


326 


3 


575 


3 


802 


4 


017 


4 


215 


8s 


1.386 


1.949 


2.376 


2.730 


3.037 


3 


310 


3 


559 


3 


786 


3 


999 


4 


196 


86 


1.379 


1-939 


2.364 


2.716 


3.022 


3 


294 


3 


543 


3 


768 


3 


981 


4 


177 


87 


1.372 


1.929 


2-352 


2. 702 


3.008 


3 


279 


3 


527 


3 


752 


3 


963 


4 


158 


88 


1.365 


1 .920 


2.340 


2.690 


2.994 


3 


265 


3 


511 


3 


734 


3 


945 


4 


139 


89 


1.358 


1 .910 


2.330 


2.678 


2.981 


3 


250 


3 


495 


3 


718 


3 


927 


4 


120 


90 


1.351 


1 .901 


2.319 


2.666 


2.967 


3 


235 


3 


479 


3 


700 


3 


909 


4 


lOI 


91 


1-345 


1.893 


2 . 309 


2.654 


2.954 


3 


221 


3 


463 


3 


684 


3 


891 


4 


082 


92 


I • 339 


1.884 


2. 298 


2.642 


2.940 


3 


206 


3 


447 


3 


666 


3 


873 


4 


064 


93 


1.333 


1.876 


2.288 


2.630 


2.927 


3 


191 


3 


432 


3 


650 


3 


855 


4 


048 


94 


1.327 


1.867 


2.278 


2 618 


2.914 


3 


177 


3 


416 


3 


634 


3 


840 


4 


032 


95 


1. 321 


1.859 


2. 267 


2.606 


2.900 


3 


162 


3 


401 


3 


618 


3 


825 


4 


016 


96 


1. 315 


1.850 


2.257 


2-594 


2.887 


3 


148 


3 


387 


3 


604 


3 


810 


4 


000 


97 


1.309 


r.842 


2. 246 


2.582 


2-873 


3 


135 


3 


373 


3 


590 


3 


795 


3 


984 


98 


1 . 303 


1.833 


2.236 


2.570 


2.862 


3 


123 


3 


360 


3 


576 


3 


780 


3 


969 


99 


1.297 


1.82s 


2. 226 


2 560 


2.851 


3 


no 


3 


347 


3 


562 


3 


765 


3 


953 


100 


1 . 291 


1. 817 


2. 217 


2-550 


2.840 


3 


098 


3 


334 


3 


548 


3 


750 


3 


937 



228 



COMPRESSED AIR PLANT 



intersection on axis A. A line joining the point on A with the 
given cu.ft. of air will cut scale No. i at the required length of 
pipe. 

4. To find the pressure at which the air must enter the pipe; 
given, maximum allowable drop, volume of air, and the diam- 
eter and length of pipe. 

Join the length of pipe with the volume of air, and note 

No 3 

2,000 







A 


No 1 




10,000^ 






0,000-i 






8,000" 






7,000 i 






6,0004 






5,000^ 






4,000 i 






3,000i 






■g 2,000^ 






a, 






fc : 






~ 



















p. 






a; 1,000- 






V. 900H 






° 800i 






■S 700 4 






^ 600-E 






a> : 






k1 500i 






400^ 






•300i 






200-^ 






100- 




A 



rl,500 



rl.OOO 
h 900 
h 800 S 
r 700 I 



r 600 S 
^ 500 "S 



r iOO g 

: u 

t> 

c. 

-- 200 I 
> 

ft 

'■< 
soo "S 

150 ^ 



No 4 



No. 3 

r-12 
-11 
-10 

1-9 
8 

7 

6 



'"' 120 


i30q 


o* „ 


116- 


CO 100 




(> 


""yo- 




80- 


^ 


70- 




60- 


.5 


50- 


"? 


40- 


to 




<3 



30- 








h 


20- 


3 




£ 


10- 


CLh 








.2 
















1-1 


0- 



No 
rO.5 
0.6 
-O.T 
-0.8 
rO.9 
rl.O 



-1.5 S 

: ^ 

- .0 

hs.o I 

I- a> 

^1.0 £ 

r a 

r5.o -^ 

-6.0 2 
F7.0 ft 

hs.o 

r9.0 
-lO.O 

4S.0 



Ll 



Join the length of pipe and the cubic 
feet of compressed airi note the intersec- 
tion with axis A; join the initial pressure 
^^'^ and the drop in pressure; note the inter- 
OQ section with aris B; join the intersections 
on A and B and from the central scale 



f 80 
h 70 



E- 60 



read the diameter.of the ^ipe 





A 








B 










u 


u 


2 






s 












a 








p. 






< 


"S 






• 


s 


£ 


\^ 




"3 


E 


-i 


--7 


• 


1 


.c 


'' 




1" 


-0 


• 






5 


ti 




\ 


b. 


0, 


.^ 









J 




\ 
\ 


3 

5 


s 


'a 






1^ 





A 








B 





Fig. 



121. 



intersection on axis A. Join this point with the pipe diameter, 
and note intersection on axis B. A line joining the point on B 
with the allowable drop will cut scale No. 4 at the required 
initial pressure. 

5. To find the pressure drop; given, initial pressure, volume 
of air, and length and diameter of pipe. 



CONVEYANCE OF COMPRESSED AIR IN PIPES 



229 



Join the pipe length with the volume of air, and note inter- 
section on axis A. Join this point with the pipe diameter, and 
note intersection on axis B. A line joining the point on B with 
the initial pressure will cut scale No. 5 at the required pressure 
drop. 

In solving six problems by the nomogram, the mean error 
was less than 0.5%. 



Note: Thlsnomogram is based on equation 
Py=:P'v' vviiere P and P"are initiai and final 
Abs. Pressure, in lbs. per sq In ,andV.V' 
the volume of free and compressed air, la 
cu ft 



^ 






14.75 -1 


r ^ 


1 






»S 


14.20- 


- 1,000 


2 

§70- 

r 50- 


150- 
100^ 

_60 j 


CQ 

u 

(B 
P. 


13.67 - 
13.16 - 
13.67 - 
12.20 - 


- 2,000 

- 3,000 

- 4,000 

- 5,000 15 





*40-! 




11.73 - 


- 6,000 W 


2 


30 i 


3 


11.30 - 


- 7,000 -9 


3 

m 

2 


20-1 


2 


10.87 - 
10.46 - 


- 8,000 -S 

- 9,000 .■§ 


5 9- 


10 i 





10.07 - 


- 10,000 ^ 


8^ 




9.70- 


- 11,000 


"o 




8 


9.34- 


- 12,000 


^ 




a 
•♦J 


8.98 - 


- 13,000 






<i 


8.65- 

8.32-1 


- 14,000 
L 15,000 



■^ 








"O 


-■ 






.a 


§ 


p. 

e 


tc 






«j 


£ 




<0 









p. 


^ 


*- 


•S 






a 


1 


— ^ 


3 


--' 


1 


1 








IS 











3 


•3 














X 







Join the cubic feet of free air with 
the altitude or atinosphfiric pressure; 
note the intersection on axisX; join 
this intersection with the absolute 
pressure of the compressed air and 
read the cubic feet of compressed 
air on the left hand scale 



(X 



Fig. 122. 

Note. — ^The nomogram, Fig. 121, uses volume of compressed 
air^ but the compressor capacity and the air consumption of 
rock-drills are usually expressed in terms of cu.ft. of free air. 
Volumes of free air at different altitudes can be converted into 
volumes of compressed air at different absolute pressures by 
means of the nomogram in Fig. 122, which is based on the equa- 



-40,000 
-30,000 

r 20,000 



-10,000 

- 9,000 

- 8,000 

- 7,000 

- 6,000 

- 5,000 

I- 4,000 
I- 3,000 



E- 2,000 2 



h 



1,000 
900 
800 
700 
600 
500 
I- 400 

=- 300 



200 



- 100 



SO 
70 
60 



230 



COMPRESSED AIR PLANT 



tion, PV = P'V (Chap. III). The error in using this nomo- 
gram is from about —0.15% to +0.5%. 

Richards* Formula, best for long pipe lines, is: 



H = 



V^L 



io,oooD% 

in which: D= diameter of pipe, inches; L = length of pipe, ft.; 
V = volume of compressed air delivered, cu.ft. per min. ; H = head 
or difference of pressure required to overcome friction and main- 
tain the flow, lbs. ; a = constant for nominal diameter of pipe. 

Values of a for Different Nominal Diameters of 
Wrought-Iron Pipe 



I 0.350 



0.500 



I4 . 



22 • 



5 0.934 

6// 
I . 000 

8^^ 1. 125 

10" 1 . 200 

12" I. 260 



3 0.730 

sh" 0.787 

4" 0.840 

0.565 
0.650 

* The values of a for i j- and li^-in. pipe are not consistent with those for other 
sizes. See foot-note on p. 223. 

By Richards' formula, the calculated losses of pressure are 
smaller, and, conversely, the volumes of air discharged are 
larger, under the same conditions, than those obtained from 
D'Arcy's formula. 

The losses of pressure in a table by F. A. Halsey show that 
the constants used by him differ materially from those given 
above. Table XIX, containing a series of random examples, 
shows that in all cases the figures from D'Arcy's formula lie 

Table XIX 



Cu.ft. Free Air 
Transmitted at 


Length of 
Pipe. Ft. 


Diameter of 
Pipe, Ins. 


Transmission Losses, Lbs. 


75 Lbs. 
(P = i5 Lbs.) 


Richards. 


D'Arcy (Cox). 


Halsey. 


1,000 


1,000 


4 


3-23 


371 


5.02 


1,000 


T,000 • 


5 


•95 


I. 17 


163 


1,000 


T,000 


6 


■35 


.46 


.64 


4,000 


5,000 


8 


592 


8.44 


13 05 


4,000 


5,000 


10 


1.78 


2.81 


4. 20 


4.000 


5,000 


12 


.68 


I .06 


1.70 



CONVEYANCE OF COMPRESSED AIR IN PIPES 



231 



between the others; hence it would appear that the results 
from this formula are sufficiently accurate for ordinary calcula- 
tions. Within certain limits, the loss of head increases with 
the square of the velocity. To obtain the best results the 
velocity of flow in main air pipes should not exceed 20 or 25 ft. 
per second (Table XX). 

Table XX*. — (Diameter of Pipe, 12 Inches) 



Velocity of Flow- 
in Ft. per Sec. 


Initial Pressure, 
Lbs. 


Final Pressure, 
Lbs. 


Per Cent of Initial 
Pressure Lost per Mile. 


25 
100 


100 
100 
100 


97.6 
go. 6 

53-8 


2.4 
9-4 

46. 2 



* Unwin. Van Nostrand's Science Series, No. io6, p. 78. 

When the initial velocity much exceeds 50 ft. per second 
the loss becomes large; but by using piping large enough to 
keep down the velocity the friction loss is almost eliminated. 
For example, in transmitting 875 cu.ft. of free air per minute 
at an initial pressure of 60 lbs., through an 8-in. pipe 7,150 ft. 
long, the average loss including leakage was 2 lbs. The velocity 
in this case was 8| ft. per second. A volume of 500 cu.ft. of 
free air per min. , at 75 lbs. gage, can be transmitted through 
1,000 ft. of 3-in. pipe with a loss of 4.1 lbs., while if a 5-in. pipe 
were used the loss would be reduced to .24 lb., the velocities 
being respectively 28 ft. and 10 ft. per second. In driving the 
Jeddo mining tunnel, at Ebervale, Luzerne Co., Penna., two 
3i-in. drills were used in each heading, with a 6-in. main, the 
maximum distance of transmission being about 10,800 ft. This 
pipe was so large in proportion to the volume of air required 
(about 230 cu.ft. free air per min.) that the velocity was only 
3I ft. per second. A calculation shows a loss of .002 lb., and 
the gages at each end of the main were found to record practically 
the same pressure. 

Due regard for economy in installation must limit the size of 
piping, the cost of which in any given case should be considered 
in relation to the cost of air compression. Diameters of 4-6 
ins. for the mains are large enough for 6-10 drills. Up to a length 



232 



COMPRESSED AIR PLANT 



of 3,000 ft. a 4-in. pipe will carry per min. 480 cu.^t. of free air 
compressed to 82 lbs., with a loss oi 2 ibs. This volume of air 
will run four 3-in. drills. Under the same conditions a 6-in. 
pipe, 5,000 ft. long, will carry 1,100 cu.ft. of free air per min., 
or enough for 10 drills in constant operation. Branch pipes 
should not be too small. For a length of, say, 100 ft. a ij-in. 
pipe is small enough for one drill, though i-in. is often used. 
While it is admissible to increase the velocity in short branches 
considerably beyond 20 ft. per second, extremes should be 
avoided. To run a 3-in. drill from a i-in. pipe 100 ft. long 
requires a velocity of flow of about 55 ft. per second, causing a 
loss of 10 lbs. (see Table XXI). 

Table XXI. — (Norwalk Iron Works Co.) 



Nominal Si 
of Pipe. 


ze 


I in. 


li 


ins. 


\\ ins. 


2 ins. 


2} ins. 




Length of 
in Ft. 


Pipe 


50 


100 


100 


300 


100 


300 


200 
84-7 


500 
53-6 


250 


600 








79.8 


23-2 


16.4 


35-2 


20.3 


63-6 


36.7 


142.0 


91.7 


< 






79.6 


33-1 


23-4 


49.7 


28.7 


89.9 


51-9 


119.6 


75-7 


200.9 


129.6 


1— 1 


Oh 




79-4 


40.4 


28.6 


61 .0 


35-2 


109. I 


63.0 


146.5 


92.7 


244.4 


157-7 


< 


K, 




79.2 


46.8 


33-1 


70-3 


40.6 


127. I 


73-4 


169. I 


107. 1 


283.2 


183. 1 


Q 


w 




79 


52.3 


37-0 


78.6 


45-4 


142.0 


82.0 


189. I 


119. 7 


317.1 


204.6 


C/3 

m 
W 
Pi 




1^ 


78.8 


57-1 


40.4 


86.1 


49-7 


155-4 


89.7 


207.0 


131. 


348.4 


224.8 


fc 




> 


78.6 


61.6 


43-6 


93-0 


53-7 


168.0 


97.0 


223.3 


141-3 


3770 


243 - 9 




w 




78.4 


65-9 


46.6 


99.2 


57-3 


179-3 


103-5 


238.7 


151-1 


399-6 


258.4 


u 

<; 
Pi . 


Q 


78.2 


70-3 


49.7 


105.4 


60.8 


190.5 


IIO.O 


252.9 


160. 1 


424.1 


273-6 


p 




H 


78.0 
77.8 
77.6 


73-7 
77.2 
80.7 


52.1 
54-6 

57-1 


no. 8 
116. 2 

121 .4 


64.0 
67. 1 

70. 1 


200. 7 
209.9 
219. 1 


115-9 
121 . 2 
126.5 


266.5 
279.2 
291-5 


168.7 
176.7 
184.5 


446.7 
469.0 
489.6 


288.6 
302.6 
315-9 




W^ 




a. 


77-4 


84.0 


59-4 


126.3 


72.9 


228.1 


131-7 


303-4 


192.0 


509 -3 


328.6 


Pi 
(—1 
< 




w 


77-2 


87.1 


61.6 


131. 1 


75-7 


236.7 


136.6 


314-4 


199.0 


528.3 


340.8 


H^ 


W 
H 


77.0 


903 


63-7 


135-4 


78.2 


245.2 


141. 6 


325-5 


206.0 


546. 5 


352.6 




H 


76.8 


92.9 


65-7 


139-8 


80.7 


252-4 


145-7 


336.1 


212. 7 


564-2 


364.0 


l-H 
Pi 

W 

Oh 


•< 

C/3 


76.6 


95-6 


67.7 


143-9 


83.1 


.259-8 


150.0 


346.2 


219. 1 


581.3 


3750 


W 


w 
p 

CO 

W 


76.4 
76. 2 
76.0 


98.4 
loi .0 
103.8 


59.6 
71-5 
73-4 


148. 1 
152. 1 
156. 1 


85-5 
87.8 
90. 1 


267.6 
274.7 
281 . 3 


154-5 
158.7 
162.4 


356.0 
365-6 

375-6 


225.3 
231-4 
2373 


597-5 
613-8 
629 3 


385-5 
396.0 
406.0 





S 


Pi 


75-8 


106.3 


75-2 


159-7 


92. 2 


288.4 


166.6 


383-9 


243-0 


644 -5 


415-8 


H 







75-6 


108.7 


76.9 


163-3 


94-3 


295-5 


170.6 


392.8 


248.6 


659.2 


425 -3 




(—1 




75-4 


III .0 


78.5 


167.0 


96.4 


301 . 7 


174.2 


401.4 


254.0 


673-8 


434-7 









75-2 


113-3 


80.1 


170.4 


98.4 


307 -9 


177.8 


409-7 


259-3 


687.8 


443-8 








75-0 


115-5 


81.7 


173-9 


100.4 


3143 


181. 5 


417-9 


264 • 5 


701 .6 


452.7 





CONVEYANCE OF COMPRESSED AIR IN PIPES 233 

Compressed-Air Piping is of wrought iron, with sleeve 
coupHngs or cast-iron flanges into which the ends of the pipe are 
expanded or screwed. Sleeve couplings are used for all except 
large sizes. The smaller sizes, to li in., are butt- welded, while 
all from i| in. up are lap- welded to insure necessary strength. 
Extra heavy piping may be had for high pressures. Wrought- 
iron spiral-seam riveted, or spiral-weld steel, tubing, sometimes 
used, is made in lengths of 20 ft. or less. For convenience 
of transport in remote regions rolled sheets in short lengths 
may be had, punched around the edges, ready for riveting, and 
packed, 4, 6 or more sheets in a bundle. 

All joints in mains and branches should be carefully made. 
The pipe may be tested from time to time by allowing the air at 
full pressure to remain in the pipe long enough to observe the 
gage. Leaks should be traced and stopped immediately; they 
are more expensive than steam leaks, because of the losses 
already suffered in compressing the air. In putting together 
screw joints see that no white lead or other cementing material 
is forced into the pipe; it would make ridges and increase the 
friction loss. Also, each length should be cleaned of all foreign 
substances which may have lodged inside. For ready inspection 
and stoppage of leaks, the pipe should, if buried, be carried in 
boxes sunk just below the surface of the ground; if underground, 
it should be supported on brackets along the side of the mine 
workings. Low points in pipe lines form " pockets " for the 
accumulation of entrained water, and should be avoided, as 
they obstruct the passage of the air. In long Hnes, where a 
uniform grade is impracticable, provision may be made near 
the end for blowing out the water at intervals, when the air 
is to be used for pumps or other stationary engines. 

For long lines expansion joints are required, especially when 
on the surface. Underground they are not often necessary, 
as the temperature is usually nearly constant, except in shafts 
or tunnels, where there may be considerable variations of tem- 
perature between summer and winter. 

As each bend or elbow in a pipe line increases resistance, 
abrupt changes in direction and sharp curves should be avoided. 



234 



COMPRESSED AIR PLANT 



For the same diameter of pipe the resistance due to a bend 
increases as the radius of the curve diminishes. In the absence 
of exact data the following table is given: 



Table XXII.— 


INORWALK Iron W 


ORKS 


Co.) 




Radius of elbow in terms of 
diameter of pipe 


5 


3 


2 


ih 


i| 


I 


3 

4 


1 
2 


Equivalent length of straight 
pipe in terms of its diameter. 


7.85 


8.24 


903 


10.36 


12. 72 


17-51 


3509 


121 . 2 



These allowances are none too large, since for steam piping 
the frictional resistance of an ordinary right-angled elbow is 
considered equivalent to that due to a length of straight pipe 
equal to 40 times its diameter. But, the usual bends in wrought- 
iron air piping are not necessarily so sharp as a standard elbow. 
When many sharp bends are permitted, the resistance may 
become very great. The matter should have special considera- 
tion in the stopes of mines, especially when timbered with square 
sets; as far as possible, the piping should be carried diagonally 
through the sets, bending the pipe itself where necessary, instead 
of using right-angled elbows. 



CHAPTER XVII 
COMPRESSED AIR ENGINES 

Compressed air may be employed as a motive power in an 
engine in two ways, vk-., at full pressure or expansively. By 
working at full pressure it is understood that the air is admitted 
to the cylinder throughout practically the entire length of stroke, 
i.e., without cutoff, and that therefore nearly a cylinderful 
of air at gage pressure is exhausted at each stroke. In this case 
the work of the air engine is roughly similar to that done in a 
non-expansive-working steam engine. Among the machines 
which use air in this way are rock-drills and simple, direct- 
acting pumps, without rotary parts. 

By the term expansive- working it is meant that the air is 
admitted to the cylinder during only a part of the stroke, and is 
then cut off and the stroke completed by the expansive force of the 
air. For operating in this way some equalizing agent, such as the 
fly-wheel, is essential, and as a rule a higher initial pressure is 
employed than when working under full pressure throughout 
the stroke. It is necessary to distinguish between complete and 
partial or incomplete expansion. When the air is used with com- 
plete expansion the operation in the cylinder is the reverse of 
adiabatic compression in a compressor, the final pressure being 
equal to that of the atmosphere. But as air does not undergo 
condensation, it follows that the lowest terminal pressure in 
the cylinder must still be suJSiciently above atmospheric pressure 
to produce a proper exhaust, and to overcome the friction of the 
engine at the end of the stroke. Hence, theoretically complete 
expansion is impracticable for simple air engines of ordinary 
design. 

Most air engines work with partial or incomplete expansion, 
the air expanding adiabatically in the latter part of the stroke. 

235 



236 COMPRESSED AIR PLANT 

The point of cutoff is such that the terminal cyhnder pressure 
exceeds the back-pressure by an amount sufficient to cause a free 
exhaust. In the conditions here set forth, no reference is made 
to the thermal changes incident upon adiabatic expansion in the 
air cyhnder. Although, in principle, compressed air is used like 
steam, both being elastic fluids, there is an essential difference in 
the results obtained, due to the reduction in temperature. In 
expanding behind the piston, a given volume of compressed air at 
a given pressure will not produce the same amount of power as 
steam under the same conditions. If two curves be constructed, 
representing the expansion of ecfual volumes of air and steam, 
from the same initial pressure down to pressures below that of 
the atmosphere, it will be seen that the steam pressure at all 
points of the stroke is considerably higher than the air pressure; 
and the expansion curve of the air reaches the atmospheric line 
sooner than the steam curve. 

Fig. 123 shows an ideal card, in which the initial pressure is 
75 lbs., and the cutoff is at I stroke. The adiabatic expansion 
curve of the air shows that the pressure is reduced to zero gage 
pressure when the air has expanded to 3 J times the initial volume, 
the mean effective pressure being 18.9 lbs. At the end of the 
stroke the pressure falls to 7 lbs. below atmospheric pressure. 
The steam curve, on the other hand, does not cut the atmospheric 
line until the expansion reaches 4 J times the initial volume, and 
the mean effective pressure is 25.2 lbs. The lower mean pressure 
of the air is due to the development of cold during its expansion. 
The operation is the reverse of compression, and the resulting loss 
of motive power is analogous to the loss of work in the compressor 
caused by the generation of heat. Just as the heat of com- 
pression reacts upon the air while being compressed in the cylin- 
der, and produces a higher tension than that due to the mere 
reduction in volume; so conversely, when expansion takes place, 
the air, which is usually at normal atmospheric temperature on 
entering the cylinder, rapidly gives up its sensible heat, and the 
cold reacting upon the expanding air reduces its pressure faster 
than that which is due to the increase in volume alone. More- 
over, this behavior of compressed air is independent of the initial 



COMPRESSED AIR ENGINES 



237 



temperature, since the resulting expansion curve would be unal- 
tered. In the case of steam the initial temperature is high, and 

LENGTH OF STROKE 



T5 



70 



60 



50 



2 40 

Z 

o 
uT 

OC 30 

CO 
OD 
UJ 

cc 
a. 

20 



10 





■■" 




~"^ 


^^ 




"~" 








~~ 




— " 


"■" 










— 










Ma« 









■" 














































































































































































































































































































































































































































































































































































\ 


























































'l\ 




























































v 




























































y 




























































\ 




























































\ 




























































\ 




























































\ 


























































































































\ 




























































\ 




























































) 






























































































































y 




























































\y 






























































V 




























































\ 


























































V 




\ 
























































\ 




N 


s. 






















































> 


V 




s 


s 






















































\ 






\ 


^A 






















































V 






^\ 




















































1 








\ 


^ 




















































\ 








\ 


s 




















































\ 










N 




















































\ 










X 


V 


















































\ 












^ 


"^ 
















































S 


s. 












^ 


















































V. 


•n, 














"^ 




^ 












































< 


































































*^ 

























































































































































































































V 


ac 


IIU 


na 


Lii 


le 


^ 


























__ 




^.^ 


^_ 




^_ 






_ 




„,, 


.,. 


1 


Jl 










m^^ 


^^ 






, 




^mm 



70 



60 



50 



40 



30 



20 



10 



Fig. 123. — Expansion Curves of Steam and Air. 

is reduced but little during expansion from ordinary working 
pressures down to atmospheric pressure. 



238 



COMPRESSED .\IR PLANT 



A similar comparison may be made for other initial pressures 
and ratios of cutott. In every case the mean effective pressure 
is higher for steam than for air. It follows that, to develop the 
same amount of power in a given cylinder and "^dth the same 
initial pressure, the cutoff must be later in the stroke -^ith air 
than with steam. 

So low are the temperatures produced by the expansion of air, 
from ordinary working pressures of 60 or 70 lbs. do'WTi to atmos- 
pheric pressure, that for a long time the expansive use of com- 
pressed air was considered impracticable. In Table XXIII 
are given the theoretical final temperatures of the exhaust air, 
in working ^-ith complete expansion, and also at full pressure 
throughout the stroke, for different ratios of initial to final 
pressure, together ^^ith the theoretical efficiencies. The initial 
temperature is taken as 68° F.* 

Table XXIII 



Ratio of 
Initial to 


Working with Complete 
expaxsiox. 


Working .a.t Flt-l Pressl're. 


Final Pressure. 


Final Tempera- 
ture, Deg. F. 


Theoretical 
Efficiency. 


Final Tempera- 
ture, Deg. F. 


Theoretical 
Efficiency. 


2 


-28.2 


•855 


- S.4 


.82 


3 


— 76.0 




806 


-34-5 




72 


4 


-106.6 




782 


-45-7 




67 


5 


-128.2 




768 


-54-4 




63 


6 


-144.4 




758 


-59-8 




60 


7 


-158.8 




751 


-63 -4 




57 


8 


-170.8 




746 


-66.1 




55 


9 


-180.6 




742 


-68.0 




53 


10 


-189.2 




7,^0 


-6q.: 




51 



In the table it is shown that by working at full pressure 
extremely low temperatures of exhaust are avoided; but the 
elficiency of this method of using compressed air is necessarily 
much below that obtained from expansive working. It is under- 
stood that the temperatures here given are theoretical and are 
never actually reached in practice. The cold produced is modi- 
fied by several causes: (i) Some heat is transmitted from the 

* ^^ Mallard. " Etude Theoretique sur les Machines a Air Comprime," p. 27. 



COMPRESSED AIR ENGINES 239 

external atmosphere through the cyHnder walls; (2) the re-com- 
pression of the clearance air at each stroke produces heat in the 
cylinder, to a degree that increases with the initial pressure and 
the clearance volume; and, (3) the presence of even a small quan- 
tity of moisture in the air tends in some degree to raise the 
cylinder temperature. 

A few brief notes will here be given concerning the elements 
of the operation of compressed-air engines, that may be con- 
sidered more or less applicable for ordinary service, viz., working 
at full pressure, with partial expansion, or with complete expan- 
sion. Isothermal expansion may be neglected, since it involves 
the application of a sufficient degree of external heat to the air 
while doing its work in the cylinder to produce a terminal tem- 
perature equal to the initial temperature. 

I. Working at Full Pressure. This mode of using compressed 
air is common for engines like pumps, operating under a constant 
resistance and not provided with fly-wheels: 
Let P' = the absolute initial pressure of the air ; 

V — the initial volume of air, at the pressure P', or K times 
the volume of i lb. of air used per unit of time; 

T' = the absolute initial temperature of the compressed air; 

T =the absolute final temperature of the air at exhaust, on 

expanding to atmospheric pressure; 
P = pressure of the air at exhaust ; 

W = foot-pounds of work done. 

From the theory of compressed air: 
R = J(Cp—C.) = 778(0.2375 -0.1689) = 53.37, where J is 
Joule's heat unit, and Cp and Cv are the specific heats of air at 
constant pressure and constant volume. 

As no work is done by the expansive force of the air originally 
produced by compression, W equals the volume of air used, V, 
multiplied by the difference between P' and P, or: 

W = V(F-P) 

KRT' 

Substituting for V its value, — ^7 — : 

^(P'-P)=S3-37^KT'(i-|) 



W 



240 COMPRESSED AIR PLANT 

2. Working with Partial Expansion. The advantages of 
using compressed air in this way may be obtained from engines 
possessing fly-wheels, provided that the cutoff be not too early in 
the stroke to avoid excessive reduction of cylinder temperature, 
or else that the air be reheated before entering the cylinder. 

In this case the values of P', V, and T' are as above. 
From the point of cutoff the air expands adiabatically dow^n to a 
terminal pressure of P^' and volume V'^ the final temperature 
in the cylinder falling to T". On exhausting, the pressure, 
volume, and temperature become P, V, and T. The work done 
is composed of three parts, Ws., 

W' =work between the point of admission and the point of 

cutoff = PV; 
W''' =work performed by expansion of the volume V from 

the point of cutoff to the end of the stroke = 778 KCp 

{T-V); 
W''' = negative work due to back-pressure = — PV. 
Taking the algebraic sum of these three quantities: 

w=pv+778Ka(r-ro-PV'' 

/ \ TT/ KRT ,, KRT 

But, asunder (i): V =—57— and V = — w7~ 

Substituting these values of V and V'^ and for R and Cp, 
their numerical values of 53.37 and 0.1689: 

W = k[s3.37T' + i3i.4(T'-T")-53-37T'(|7)] 

= 53.37k[t' + 2.46(T'-T")-T'p 

3. Working with Complete Expansion. In the theoretical 
card. Fig. 124, are shown the relations of the compression and 
expansion lines, the shaded portion representing the useful work 
done by the complete expansion of cold air in a motor cyHnder. 
When the expansion is adiabatic, the same relations exist between 



COMPRESSED AIR ENGINES 



241 



pressures, volumes, and temperatures as were set forth in the dis- 
cussion of adiabatic compression, viz: 

p/ /SJ \n=lA06 /'T'f\n-1 



— . - " 

P VV7 \T/ 



0.29 




70 a 



a< 

P. 



S 



CM 



Vols, in Cu. Ft. 
Fic. 124. 



The theoretical work done by complete adiabatic expansion 
may be expressed by a formula like that employed for compres- 
sion, but with an inversion of certain of the quantities, thus: 

/ p \ It — I' 



W = 



n 
n—i 



PV 



I — 



P' 



in which W = theoretical foot-pounds of work done by the ex- 
pansion to atmospheric pressure of i lb. (13. i cu.ft.) of free air. 
Substituting the values of the constants, and for working at sea- 
level : 

0.29" 



w 



3.463Xi44Xi4.7Xi3.iX[i-(^^j 



F= 96,029 



I — 



14-7 
P' 



0.29 



242 



COMPRESSED AIR PLANT 



For example, if P' be 40 lbs. gage pressure: 



W= 96,029 



I — 



14-7 
54.7 



0.29 



= 30,440 ft. lbs., or 2,323 ft.-lbs. per cu. 



ft. of free air. 

Actual Work Done. In the above expressions no account is 
taken of the friction of moving parts of the motor engine, or loss 
of work caused by leakage. In determining the actual work, the 
general case will be where a cutoff is employed. The relations 
between initial and terminal pressures and temperatures, for 
different ratios of expansion in a motor-engine cylinder, are 
shown in Table XXIV,* the points of cutoff, in tenths of the 
cyUnder stroke, being given in the first column. 

Table XXIV 

Theoretical Ratios of Pressures and Temperatures Due 
TO THE Expansion of Compressed Air in a Motor 
Cylinder 







Ratio of 


Ratio of 




Ratio of 
Initial to 
Final Abso- 
lute Tem- 
perature, 
Due to Ex- 


Ratio of 






Mean to 


Mean to Ri 


itio of 


Initial to 




Ratio of 


Total Abso- 


Total Abso- In 


tial to 


Final Abso- 


Cutoff. 


Expansion 


lute Pres- 


lute Pres- 1 


^inal 


lute Pres- 




= I -7- Cutoff. 


sure, for 
Entire 


sure. During To 
Expansion 1 


mpera- 
,ure. 


sure for 
Ratio of 






Stroke. 


Only. 




pansion 
Only. 


Expansion. 


0. 10 


10.00 


0.249 


0.166 


391 


0.513 


0.039 


•15 


6.67 




348 


■^33 


460 




578 




069 


.20 


5.00 




436 


•295 


518 




627 




104 


•25 


4.00 




515 


•353 


568 




669 




142 


•30 


3-33 




585 


.408 


612 




705 




184 


•35 


2.86 




647 


.460 


652 




737 




228 


.40 


2.50 




706 


.510 


688 




767 




275 


•45 


2. 22 




757 


•558 


722 




794 




325 


•50 


2.00 




802 


.604 


754 




818 




378 


•55 


1. 81 




842 


.649 


784 




841 




433 


.60 


1 .67 




877 


.692 


812 




862 




487 


•65 


1-54 




907 


■734 


839 




882 




545 


.70 


1-43 




932 


•774 - 


86s 




902 




605 


•75 


1-33 




954 


.814 


889 




920 




667 



* This table, as well as Table XXV, is taken in part from those used by G. D. 
Hiscox, in " Compressed Air, its Production, Uses and Application," 1901, p. 202. 



COMPRESSED AIR ENGINES 



243 



The quantities in Table XXIV must be further corrected for 
piston clearance and the lost volume represented by the air ports 
and passages of the cylinder, because part of the air expands into 
these clearance spaces. Therefore, the actual effect of the cutoff, 
in any given case, is found by dividing the sum of the cutoff 
plus clearance, by the cylinder volume plus clearance. For 
example, if the stroke be io,with a cutoff of A, and clearance 
of 6%, the actual volume of the cylinder, including clearance, 
will be: (10X.06) + 10 = 10.6. Then the sum of the cutoff plus 
the clearance is 4+.6 = 4.6, and the working cutoff becomes 
4.6 -^ 10.6 = 0.434. In this way Table XXV has been constructed, 
for use in connection with Table XXIV. It shows the actual 
cutoff corresponding to the different nominal points of cutoff, for 
the percentages of piston clearance named at the top of the 
columns. 

Table XXV 

Actual Cut-off Due to Clearance, for the Nominal Cut- 
offs IN Column i 



Nominal 


Percentage of Clearance. 


Cutoff. 


















•03 


.04 


• OS 


.06 


.07 


.08 


.10 


. 10 


126 


135 


143 


0.151 


0.159 


167 


182 


•15 


175 


184 


191 




198 




206 


213 


227 


. 20 


223 


231 


238 




245 




252 


259 


273 


•25 


272 


279 


286 




293 




299 


305 


318 


•30 


320 


327 


333 




340 




346 


352 


364 


•35 


368 


376 


380 




387 




392 


398 


409 


.40 


417 


423 


429 




434 




439 


444 


455 


45 


465 


471 


477 




481 




486 


490 


500 


•50 


514 


519 


524 




528 




533 


537 


546 


•55 


564 


568 


571 




576 




580 


585 


591 


.60 


612 


615 


619 




623 




626 


630 


637 


•65 


660 


664 


667 




670 




673 


676 


682 


.70 


709 


711 


714 




717 




720 


722 


727 


•75 


758 


760 


762 




764 




766 


768 


772 



The theoretical terminal cylinder pressure resulting from 
adiabatic expansion may be expressed by : 



244 COMPRESSED AIR PLANT 

P' . . I 

P^rj^^ — VAn which C = ratio of expansion = — r— — 7 — - — ^ (see 
C^ *^^ ^ point of cut-off ^ 

column 2, Table XXIV). 

For example, for a cutoff at ^ stroke and 65 lbs. gage pressure, 
the terminal pressure (above atmospheric pressure) will be: 

65 + 14.7 * ,1 

-^;^Ii5i — 14.7 = 7. 2 lbs. 

The volume corresponding to the nominal cutoff is increased 
by the clearance, and adds to the mean pressure. Thus, in the 
above example, assuming the clearance to be 6%, the actual 
cutoff (Table XXV) is increased from 0.4 to 0.434, of which the 

ratio C is = 2.3. From Table XXIV, cojumn 7, the ratio 

•434 

of initial to terminal pressure, corresponding to the actual cut- 
off of 0.434, is (by interpolation) .31; whence: (79.7X0.31) — 
14.7 = 10 lbs. terminal pressure. 

Cylinder Volume Required for a Given Power. The work 
per stroke is found by dividing the foot-pounds of work to be 
done per minute by twice the number of revolutions of the engine 
(which would be determined for any given size of engine by 
the ordinary empiric rules of practice). This is substituted, 
with the initial and final pressures, in the formula for working with 
full pressure, partial or complete expansion, as the case may be, 
which is then solved for the initial volume, V, of compressed 
air used per stroke. To the theoretical cylinder volume thus 
found, the allowance for piston clearance is added, according to 
the type of engine. The proper proportion between stroke and 
diameter of cylinder is finally determined. 

The volumes of free air per minute, required for an air engine, 
per indicated horse-power and for different ratios of cutoff, are 
shown in Table XXVI, by F. C. Weber.* The figures given in 
this table do not include the volume corresponding to piston 
clearance, which may be found as already shown. 

* Compressed Air, Oct., 1896, p. 117. 



COMPRESSED AIR ENGINES 



245 



Table XXVI 

Cubic Feet of Free Air per Minute Used in Motor Engine, 

Per I.H.P. 



Point 


Gage Pressures, Lbs. 


of 


















< 




Cutoff 


30 


40 


50 


60 


70 


80 


90 


100 


no 


125 


I 


23-3 


21.3 


20. 2 


19.4 


18.8 


18.42 


18.10 


17.8 


17.62 


17.40 


3 
4 


i8.7 


17. 1 


16. 1 


1547 


150 


14.6 


14-35 


14-15 


13.98 


13-78 


2 
3 


17-85 


16. 2 


15-2 


145 


14.2 


13-75 


13-47 


13.28 


13.08 


12.90 


1 
2 


16.4 


145 


135 


12.8 


12.3 


11-93 


II. 7 


11.48 


11-30 


II . 10 


1 
3 


175 


152 


12.9 


11.85 


II . 26 


10.8 


10.5 


10. 21 


10.02 


9-78 


1 
4 


20.6 


15-6 


13-4 


133 


-1 1 . 40 


10. 72 


10.31 


10. 


9-75 


9.42 



In this table the air is supposed to be used without reheating, 
and at an initial temperature of 60° F. Reheating will reduce the 

T2 

volume of air proportionally to the ratio =r-, where T2 =459^+60° 

I3 

= 519° F., or absolute temperature; and T3 = 459° plus the tem- 
perature of the reheated air on entering the motor cylinder. 
Thus, if the air be reheated to 200° F., the above ratio becomes 

o 

^ — 5 = 0.787, by which decimal the volume of air as found in 

the table must be multiplied. 

So far as mine service is concerned, it has been customary 
to consider compressed air almost exclusively as an agent for the 
operation of rock-drills, and in view of its preponderating applica- 
tion to this use its adaptability under proper conditions to the 
driving of other machines and engines is sometimes overlooked. 
Of late years, however, with improved methods of compression 
and reheating, attention has been given to employing compressed 
air for a greater variety of service; not only underground, but 
for certain portions of the surface plant of mines as well. Aside 
from cases where the disposal of exhaust steam would be trouble- 
some, the question is largely one of comparative loss in trans- 
mission and the power cost of the air. 

Although not strictly in place in this chapter, reference may 
be made to what has been called the " two-pipe system " or 



246 COMPRESSED AIR PLANT 

" high-range compressed-air transmission," introduced some 
years ago by Charles Cummings.* 

The machine or engine using the air makes in effect a closed 
circuit with the compressor. After the air has done its work in 
the motor cyhnder, it is returned to the compressor at the pressure 
of the exhaust, through a second line of piping. The return pipe 
connects with a closed chamber at the compressor, in which the 
inlet valves are placed, thus enabling the compressor to begin its 
stroke \\dth the cylinder filled under a considerable initial pressure. 
Then, after raising the pressure to the original point, the com- 
pressor delivers the air into the main, to be used again by the air 
engine. The actual working pressure of the air engine is, there- 
fore, the difference between the pressures in the delivery and 
return pipes. Barring leakage, the same air is thus used over and 
over, the intention being that the compressor shall put back into 
the air kept in circulation the power expended in the motor- 
engine cylinder. 

Though the compressor itself is not materially different from 
the ordinary forms, the two-pipe system requires a rather com- 
plicated arrangement of piping and valves for charging the 
apparatus with air at the working pressure adopted, and for 
governing the speed and output according to the rate of con- 
sumption of air.f The advantages of the system are: a higher 
efficiency than is obtained from moderate- size compressors of the 
usual types, and less trouble from freezing at the motor engine 
by reason of the relative dryness of the air due to its higher 
tension. The efficiency increases with the pressure employed. 
In using compressed air without reheating the two-pipe system 
is superior in principle to the ordinary mode of operating com- 
pressed-air plant. But because of the greater first cost its 
advantages disappear when reheating can be adopted, and the 
single-pipe system is then found to be preferable. 

The two-pipe system is best suited for machines working 

* Patent No. 456,941 was issued to Mr. Cummings in 1891. 

t A detailed illustrated description is given by Frank Richards in American 
Machinist, April 28, 1898, p. 23. See also Compressed Air Magazine, Oct., 1907, 
P- 4599- 



COMPRESSED AIR ENGINES 247 

at full pressure throughout the stroke, such as machine drills or 
simple, direct-acting pumps. When the motor works expan- 
sively the pulsations become objectionable, as a regular flow of 
air is not maintained in the return pipe. Under these con- 
ditions the inertia and friction of high-pressure air in long pipe 
lines becomes noticeable and disadvantageous. 

As the length of air pipe required for this system is doubled, 
not only may the first cost of the pipe go far toward offsetting the 
greater efficiency but, with at least twice as many joints in the 
pipe lines, the chances of loss from leakage are increased. And 
if very high pressures be used (pressures of several hundred 
pounds have been proposed), not only must the piping itself be 
heavier and more expensive, but the proportionate power loss 
from leakage is greater. For moderate distances, however, and 
when working at full pressure under the proper conditions, the 
foregoing disadvantages may be more than counterbalanced by 
the superior efficiency of the system. Though not yet in general 
use, the two-pipe system is said to have given satisfaction at 
several mines in New Mexico, Colorado, and California,* and in 
1905 was proposed for use in the Johannesburg gold district. 
Some prominence is here given to the system because of its 
novel features and the probability that it may be found useful, if 
its disadvantages can be overcome. In a paper by H. C. Behr, 
published in 1905 in the Transactions of the Mechanical Engineers^ 
Association of the Witwatersrand, the Cummings system is treated 
at length, with a discussion of its advantages for air-driven 
pumps. 

Compressed- Air Hoists. In a few mines, compressed-air 
engines of considerable size are used in deep shafts, where the 
hoisting is in two stages. The usual case is where an incHned 
shaft is sunk in the vein, or in the foot wall rock, to a point of 
intersection with a vertical shaft. The compressed-air hoist 
at the head of the incline delivers ore into pockets at the foot 
of the vertical shaft, from which the main steam or electric 
hoist raises it to the surface. This plan is followed at several 
mines on the Witwatersrand, South Africa, as at one of the 

* A. E. Chodzko, Modern Machinery (Chicago), Jan., 1899, p. 11. 



248 



COMPRESSED AIR PLANT 



Simmer Deep Shafts. For small-scale work standard geared 
hoists are used, usually without trouble from freezing of moisture, 
because in intermittent work the cyHnders have time to regain 
normal temperature. 

For heavy work the cylinders should be designed for expansive 
use of the compressed air. The clearance volume is thus reduced, 
and larger admission and exhaust ports are required, because 
at the same pressure the density of air exceeds that of steam. 
Loss in efficiency due to incomplete expansion can be reduced 
by compounding the cylinders, and, if possible, reheating the 
air. With a cutoff in the high-pressure cylinder at 0.9 stroke 
(the minimum practicable starting cutoff) , and reheating between 
the cylinders to the initial temperature, the loss is about one-half 
that of a simple cylinder, or a saving of 25% of the energy in 
the air entering the high-pressure cylinder. At the Miami 
mine, Ariz., the reheating temperature is 350^-375° F.; at the 
Anaconda mine, 25o°-35o° F. The volume of air required and 
the results with different cutoffs are discussed on previous pages 
(see also Table XXVII). 



Table XXVII 
Volume of Free Air (60 Lbs. Gage) for Duplex Hoists 



Diameter 

of 

Cylinder, 

Ins. 


Strokes, 
Ins. 


Revs, per 

Min. 


Normal 
H.P. 


Actual 
H.P. 


Weight 

Lifted, 

Single 

Rope, Lbs. 


Free Air 

per Min., 

Cu.ft. 


5- 


6 


200 


6 


II. 8 


1,000 


300 


5 




8 


160 


8 


12.6 


1,650 


320 


6 


25 


8 


160 


12 


19.8 


2,500 


500 


7 




10 


125 


20 


24.2 


3,500 


604 


8 


25 


10 


125 


30 


33-6 


6,000 


680 


8 


50 


12 


no 


40 


37-8 


8,000 


952 


10 




12 


no 


50 


52.4 


10,000 


1,320 


12 


25 


15 
18 


100 


75 
100 


89.2 

125- 




2,250 
3,174 


14 


90 









Small underground compressed-air hoists, portable or semi- 
portable, are furnished by several makers for sinking shafts 
from level to level, sinking winzes, raising timbers into position 



COMPRESSED AIR ENGINES 249 

in stopes, temporary haulage of cars, and other Hght work. They 
are mounted on a drill column or bar, or can be bolted to a 
timber. Following are examples: 

The Inger soil-Rand "Little Tugger " (Fig. 125) is made 
in two sizes: '' I-H," for 300 ft. of i^-in. or 500 ft. of i-in. 
wire rope, to raise a i,ooo-lb. load, and " H-H," for 200 'ft. of 
|-in. manila rope, to raise 600 lbs. It will run on compressed air 
or steam. The engine is of the square piston type, giving 4 
impulses per min., so that there is no dead center. With air at 






"" ■ l£V£R I 
FOR ENG/NS 



Fig. 125. — IngersoU-Rand "Little Tugger" Hoist, Type " I-H." 

80 lbs., a load of 1,000 lbs. can be raised at a speed of 85 ft. per 
min. ( = about 2.5 H.P.). 

The Leadville '^ column-hoist " (Mine and Smelter Supply 
Co.) has a 2-H.P. engine. Other designs are the Holman, 
made in sizes of 2, 4 and 6 H.P., and the portable hoist, for 
column mounting, of the Chicago Pneumatic Tool Co., 2 H.P., 
lifting capacity, 650 lbs., at 90 ft. per min.; 200 ft. of i^-in. 
rope; weight, without rope, 300 lbs. 

Anaconda Copper Go's Compressed-Air Hoisting Plant.* 

* For further details, see papers by: B. V. Nordberg, Trnw^. A.I.M.E., Vol. 
XLVI, p. 826; and Eng. 6" Min. Jour., May 18 and 25, 1Q12. See also: K. A. 
Pauly, Trans. A.I.M.E., Vol. XLII, p. 533; R. R. Seeber, Eng. df Min. Jour., 
Aug. 3, 191 2, p. 197, and T. T. Read, Min. df Sci. Press, Nov. 2, 191 2, p. 554. 



250 COMPRESSED AIR PLANT 

From 1911 to 1913, a number of the main Anaconda hoists were 
changed to compressed-air drive by the Nordberg Mfg. Co. 
Most of the shafts range in depth from about 2,000 to 2,500 ft. 
The great variation in power required in deep hoisting, and the 
high peak loads during the period of acceleration, suggested 
the use of compressed air, in connection with an extensive 
storage system. At one double-compartment shaft, where large 
quantities of waste rock for filling were being lowered from 
the surface, the indicated horsepower ranged from —1,600 to 
+ 2 ,300. An elaborate series of tests on the former steam -hoisting 
plants led to the following plan : 

A. An electrically operated compressor plant, with a capacity 
of 7,650 cu.ft. of free air per min. compressed to 90 lbs., the 
electric current being generated by water power, 150 miles from 
Butte. 

B. Air storage of large capacity. This is connected with the 
air piping for the rock-drills, so that, in periods when little or 
no air is used by the hoists, all the air could be turned into the 
drill mains. To take care of the peak loads, each hoist has 
several air receivers, aggregating about 8,000 cu.ft. capacity. 
Maximum air consumption of the largest hoists is 60 cu.ft. of 
compressed air per second, for a period of 5 seconds; the heaviest 
unbalanced load requires an average of 40 cu.ft. of compressed 
air per second for about i min., or a total of 2,400 cu.ft. The 
pipe lines are proportioned for a pressure drop of 5 lbs. At 
times when several large hoists are running simultaneously 
there is a total draft of 2,600 cu.ft. compressed air per min., 
or 22,130 cu.ft. free air (atmospheric pressure at Butte being 
12 lbs.). The combined maximum consumption of the entire 
27 hoists would be 37,200 cu.ft. free air per min. (average, 
27,660), the highest peak representing a volume of 44,000 cu.ft. 
As this measures the minimum required storage capacity to 
equalize the load on the system, to provide for it, and allowing 
10% pressure drop, the total receiver capacity would be 440,000 
cu.ft., which is prohibitively large. 

A hydrostatic storage plant was therefore installed, con- 
sisting of a number of connected air receivers of 66,000 cu.ft. total 



COMPRESSED AIR ENGINES 251 

capacity. About 210 ft. higher than the receivers (correspond- 
ing to a pressure of 91.15 lbs. per sq. in.) is an open water tank, 
100 ft. diameter, with a pipe leading from its bottom to the 
receivers. When 66,000 cu.ft. of compressed air have been 
delivered into the receivers, the water in them is displaced 
and forced into the upper tank. Hence, the hoists are supplied 
with air at constant pressure of 91.15 lbs., less the pressure 
drop due to pipe friction. 

C. Air reheaters are an important feature of the plant, 
the heating medium being steam at 200 lbs. pressure; reheating 
temperature, 300° F.* Coal used for reheating is J lb. per 
H.P. hour. 

D. The steam hoisting engines, ranging in sizes from 28 ins. 
by 48 ins. to 34 ins. by 72 ins., were fitted with new cylinders, 
with special valve gearing, t 

The advantages of the system are: (a) the existing hoists 
were retained, with comparatively few changes ; (b) the engines 
may be run by steam if necessary; (c) the air storage is sufficient 
to operate each hoist for about 20 min., in case of stoppage of 
the electric power, due to thunder storms, etc.; (d) hoisting 
capacity can be increased by increasing speed, and the same 
engine will hoist from a greater depth by increasing the com- 
pressor capacity; (e) part of the energy liberated during the 
period of retardation, or when waste rock is being lowered, is 
returned to the power system; (/") peak loads on the power 
plant are eliminated. 

Approximate efficiency of the main hoists : 

Electric motor efficiency 95% 

Total efficiency of the compressors 74 

Efficiency of hoisting engines 50 

Total efficiency of plant, without reheating 35.4 

Total efficiency of plant, with reheating ... 53 

Other Compressed-Air Hoists have been installed at several 
mines. 

* For details see Trans. A.I.M.E., Vol. XL VI, p. 857. 

t Ibid., p. 860. See also Peele's Mining Engineers' Handbook, Sec. 12, Art. 10, 



252 COMPRESSED AIR PLANT 

I. Miami Copper Co., Ariz. A 20 in. by 48 in. Nordberg 
geared hoist, with lo-ft. drums. Air at 80 lbs. is reheated to 
370° F.; consumption, 2,275 cu.ft. free air per min., for depth of 
675 ft. Rope speed, 750 ft. Capacity, 287 tons per hour. A 
rough test showed the net shaft H.P. to be about 60% of the 
I.H.P. of compressor Cost, f.o.b. Milwaukee, $22,500. 

II. FrankUn Junior mine, ^Nlich., has a combined air and 
steam hoist.* The lo-ton empty skip is lowered unbalanced, 
and compresses air back into 3 steam and air storage receivers, 
each 10 ft. diam. by 32 ft. long. Air cylinders, 36 ins. 
diam., are tandem to 46-in. steam cylinders; stroke, 72 ins.; 
Corhss valve gear. Drum, 15 ft. diam., holds 5,130 ft. of if -in. 
rope. Two 200-H.P. boilers are used. When hoisting, the ends 
of the air cylinders are by-passed, and no air is compressed; 
when lowering, the steam-cylinder exhaust valves are open, 
and the air compressed in the air cylinders is discharged into 
the receivers. A reducing valve keeps the receiver pressure at 
75 lbs. The skip is lowered with the receivers full of steam at 
75 lbs.; the compressed air raises the pressure to 95 lbs., which 
is then available to start hoisting. When the pressure falls 
below 75 lbs., steam is admitted to complete the hoist. 

III. Copper Queen mine, Ariz., has 3-direct-acting and 2 
geared hoists operated by compressed air. The larger direct- 
acting hoists use 1,400-1,600 cu.ft. free air per shaft H.P. 

* R. H. Corbett, Eng. ca' Min. Jour., Sept. 21, 1912, p. 553; see also Power, 
Nov. 7, igi6. 



CHAPTER XVIII 

FREEZING OF MOISTURE DEPOSITED FROM 
COMPRESSED AIR 

Reference has been made in a former chapter to the trouble 
sometimes caused by the congelation of the moisture carried 
. in compressed air when deposited in the transmission pipes or in 
the ports and exhaust passages of the machine using the air. 
The presence of moisture in compressed air must be accepted as 
an unavoidable condition. Existing in the atmosphere at all 
times in greater or less quantity, when air is compressed the 
moisture is carried with it. A part of the water is deposited in 
the air receiver, but a considerable quantity still remains and will 
be brought into evidence when the proper conditions occur. 

The capacity of air for moisture depends primarily upon its 
temperature. Under ordinary atmospheric conditions i,ooo 
cu.ft. of air contain about i lb. of water. When its volume is 
reduced in the compressor cylinder, the increase of heat which 
takes place augments its moisture-carrying capacity. Any 
subsequent decrease in temperature reduces this capacity, and 
if the air be saturated the excess of moisture is deposited. 
Volume for volume, the capacity of air for moisture is independ- 
ent of its pressure or density. That is, at the same temperature, 
a cubic foot of air at atmospheric pressure will hold in suspension 
the same weight of water as a cubic foot at loo lbs. pressure. 
But this must not be misunderstood. If a certain volume of sat- 
urated atmospheric air be compressed isothermally, say to iV of 
its original volume, its water capacity is also reduced to to , and 
To of the water originally present in the air is deposited. There- 
fore, while the capacity for carrying moisture of a given vol- 
ume of air varies with the temperature, it must change also with 
any increase or decrease of pressure which changes its volume. 

253 



254 COMPRESSED AIR PLANT 

Causes of Freezing. Certain conditions are required to cause 
freezing of compressed air: deposited moisture must be present, 
and it must be subjected to a temperature below the freezing- 
point. So long as the temperature does not fall low enough, the 
presence of moisture can do no harm. Although one of the 
recognized functions of the air receiver is to permit the deposition 
of water before the air passes into the pipes, still, unless the 
receiver be extremely large, the air leaves it warm — usually even 
quite hot — and therefore carries with it considerable moisture. 
In the case of wet compressors, unless liberal sprays are used to 
attain effective cooHng, the air is apt to contain more moisture 
than that from dry compressors. A well-designed injection com- 
pressor, however, not too small for its work and therefore running 
at a moderate speed, will deliver cool air which will not give 
trouble from freezing. The air having attained nearly normal 
temperature before entering the pipe-Hne, its moisture-carrying 
capacity undergoes but little further reduction while passing 
through the pipe, and only a small amount of additional deposi- 
tion takes place. With dry compression the percentage of 
humidity of the intake air, and the temperature at discharge, 
determine the quantity of water carried out of the cyHnder. The 
humidity, in turn, varies with the weather. Changes in the 
weather may quickly be followed by variations in the quantity 
of moisture deposited in the receivei and pipe-line. When the 
air is finally expanded in doing its work in the air engine, intense 
cold is produced as the pressure falls, and the latent heat of 
compression is absorbed. It is here that the moisture carried 
with the air into the pipes makes its appearance as frost and 
causes trouble. Watery vapor itself, depositing a Hght, snow- 
like frost, does not tend to clog the air passages and ports as 
much as entrained water in a finely divided state, which will 
gradually form accumulations of solid ice and choke the exhaust 
wholly or in part. 

Prevention of Freezing. The difficulties which may arise 
from the conditions just outlined are apt to be exaggerated. 
That freezing not infrequently occurs is true, but with a properly 
designed and arranged plant it may easily be avoided. Two 



FREEZING OF MOISTURE 255 

things require attention: first, the air should be caused to drop its 
moisture as completely as possible before entering the main; 
second, provision should be made for draining off what deposited 
moisture remains in the pipe-line, before the air passes to the 
machine in which it is to be used. Although this is a pimple 
matter, the means for accompHshing it are often neglected. 
Considerable quantities of water may collect in low places in the 
pipe-line and, if not blown out at intervals, will be carried into 
the ports, cylinders, and exhaust passage of the air machine and 
there freeze. 

Granting that the air leaves the receiver near the compressor 
practically saturated and still warm, it is evident that a great 
improvement in working conditions may be realized by intro- 
ducing a second receiver as close as possible to the machines 
using the air. In mining the second receiver is, of course, placed 
underground.* Before reaching it, the temperature of the air 
will have become normal, and the entrained moisture from the 
pipe-line may readily be trapped and drawn off. It may be 
remarked that automatic water-traps are preferable to valves or 
cocks for getting rid of the water. As a rule, when the comx- 
pressed air is to be used expansively, a special after cooler should 
be introduced, placed as close as possible to the compressor. In 
any case, the receiver should be of ample size to insure the deposi- 
tion of the moisture. The advantages of reheating the air 
before use will be taken up later. 

Influence upon Freezing of High Pressures in Transmission. 
The statements made in the first part of this chapter suggest an 
important consideration, viz., in transmitting power by air at a 
high pressure there is less liability to trouble from freezing than 
when low pressures are employed, provided that the length of 
pipe-line is sufficient to allow the air to be completely cooled and 
drained of its water while still under high pressure. At a low 
pressure a greater volume of air is required to furnish a given 
amount of power than when at a high pressure. More moisture 
must, therefore, be dealt with, and at the low pressure it cannot 
be so thoroughly separated before the air is used. Suppose the 

* See Chapter XI, 



256 COMPRESSED AIR PLANT 

transmission to be at a high pressure, and .through a pipe long 
enough to allow the air to reach normal temperature. If the 
deposited moisture be drained away while the air is at its maxi- 
mum pressure; then, if the air be subsequ'ently expanded down 
to a lower pressure suitable for working (with a corresponding 
increase of volume) and allowed to regain its normal tempera- 
ture, the percentage of moisture will be reduced, so that the air 
may be relatively very dry. \A^en finally used in the air engine 
there mil not be enough moisture present to cause troublesome 
freezing. 

Deposition of Moisture by Reducing Pressure. Still another 
mode of minimizing trouble from freezing is to reduce the 
pressure of the air before it enters the cylinder of the air engine. 
The means by which this is accomplished and the results obtained 
may be illustrated by an example. 

At the Drummond Colliery, Nova Scotia, for running an 
underground pump by compressed air two receivers are used, 
one near the pump, and another 300 ft. farther back on the pipe- 
line. The air pressure in the main from the surface is 85 lbs., 
and as the proportions of the cylinders of this particular pump 
are such that so high a pressure was unnecessary a reducing 
valve was put in the pipe just before reaching the first receiver. 
By this valve the air is wire-drawn to reduce the pressure to 
45 bs., which results in a deposition of nearly one-half the 
entrained water, in addition to that already deposited in the 
pipes. It is found that more moisture collects in the first than 
in the second receiver, and by this device the serious difficulty 
previously encountered from freezing at the pump has been 
entirely overcome.* The temperature lost by the reduction of 
pressure to 45 lbs. is regained before the air reaches the pump. 

Protection of Surface Piping. What precedes refers only to 
the freezing produced by internal reduction of temperature, 
acting on the moisture carried in the air. In using compressed 

* This information has been kindly furnished by Charles Fergie, superin- 
tendent of the Drummond Colliery. See also Mr. Fergie's article on the subject, 
in Transactions Canadian Mining Inst., 1806, of which an abstract was published in 
the Colliery Guardian, October 30, 1896, p. 821. 



FREEZING OF MOISTURE 257 

air, even for mining purposes, it often becomes necessary to 
carry lines of air pipe considerable distances on the surface. 
To prevent condensation and freezing of the moisture in winter 
by external cold, all surface piping must be protected. If 
exposed to temperatures below the freezing-point, the inside 
of the pipe will become coated with ice and its effective 'cross- 
section reduced. A serious diminution of area may thus be 
caused at low points in the pipe-line, where water tends to col- 
lect; or the pipe may even be frozen solid in such places by the 
gradual accumulation of ice. Underground the temperature 
is rarely, if ever, low enough to render any protection necessary, 
except in cold, down-cast shafts, or in tunnels in which there is a 
strong inward draught. 

Some time ago, at one of the Butte copper mines, a simple and 
inexpensive device was employed to prevent the freezing of mois- 
ture in a long line of surface piping. The air main of a large 
compressor plant was carried on the surface some hundreds of 
feet before reaching the shaft. During the winter months it was 
at times difficult to get sufficient air pressure in the mine because 
of the partial choking up of the pipe. As the volume of com- 
pressed air was too large to be dealt with by the ordinary receiver, 
a series of old tubular boilers were placed close to the compressor 
house. The hot air, at 80 lbs. gage pressure, in passing through 
these boilers, from one to another, was cooled down practically 
to >atmospheric temperature and as a consequence a large part 
of its moisture was deposited. It was found that discarded 
tubular boilers, when strong enough, were well suited to this 
purpose, because of the large surface presented to the cold 
outside air; especially when they are set horizontally, so that 
there is a free circulation of air through the tubes. A blower 
might be used for the same purpose in a warm climate, or the 
boilers submerged in cold water. This effectual remedy is 
worthy of adoption where the conditions are similar. 



CIL\PTER XIX 

REHEATING COMPRESSED AIR 

After the warm compressed air enters the transmission line 
its temperature is quickly reduced to that of the surrounding 
atmosphere. The facihty ^^ith which the heat of compression is 
given up suggests the gain that may be effected by reheating the 
air when it reaches the place where it is to be utilized. By 
reheating an added volume of air is obtained at a lower power 
cost than if it were produced by a compressor. This is shown 
by comparing the heat units required to produce a given volume 
of air at a given pressure in a compressor cyHnder vdth the heat 
units required to accomphsh the same result by expanding the 
air by the direct appHcation of heat. 

Appliances for, and Results of Reheating. The most 
important methods of reheating are: (i) the air to be heated 
is passed through a cast-iron chamber or coil of pipe, exposed 
to a fire or current of hot gases or steam; (2) heat is added 
within the body of air itself, by the combustion of fuel, the 
injection of steam or hot water, or placing in the air pipe an 
electric-resistance coil. The first method is preferable from a 
mechanical standpoint and is the most efficient. The others 
may hz employed where the ordinary burning of fuel is not 
admissible. 

The follo\\ing calculation * shows the theoretical results of 
reheating: 

Weight of I cu.ft. of steam, at 75 lbs. gage = 0.2089 lb. 

Total heat units in i lb. of steam, at 75 lbs., produced from 
water at 60° F. = 1151. 

Total heat units in i cu.ft. of steam at 75 lbs. = 1151 X 
0.2089 = 240.44. 

* Frank Richards, " Compressed Air," p. 158. 
258 



REHEATING COMPRESSED AIR^ 259 

To produce by compression in a steam-actuated air com- 
pressor I cu.ft. of compressed air at 75 lbs. gage and 60° F., 
about 2 cu.ft. of steam at the same pressure are required,* 
making the thermal cost of i cu.ft. of compressed air, at the 
above temperature and pressure, 240.44X2=480.88 heat units. 
The air is here supposed to have lost its heat of compression by 
being stored or transmitted to a distance, so that the 480.88 heat 
units represent its cost at the motor where it is used. 

Result of reheating : 

Weight of I cu.ft. of air at 75 lbs. and 60° F. =0.456 lb. 

Units of heat required to double the volume of i lb. of free 
air at 60° F. = 1 23.84. 

Units of heat required to double the volume of i cu.ft. of 
compressed air at the above temperature and pressure = 123.84 X 
0.456 = 56.47. 

Comparing the thermal cost of i cu.ft. of air compressed 
in a cylinder with that of i cu.ft. obtained by reheating: 

480.88: 56.47: : i: 0.1174 

that is, the cost in heat units of a volume of air produced by 
reheating is less than | of that required to produce the same 
volume by compression. 

This theoretical result cannot be attained in practice. To 
effect such a saving a perfect r cheater would be necessary, and 
the air after reheating would have to pass directly into the 
motor cylinder. A farther conveyance of the air in pipes for 
even a short distance rapidly lov/ers its temperature and pressure. 

Reheating Temperatures. At constant pressure the volume 
of air is proportional to its absolute temperature, or 459° F. 
plus the sensible temperature above zero. The absolute tem- 
perature of air at 70° F. is 459 + 70 = 529°. In doubling the 
volume by the application of heat the absolute temperature 
becomes 1058°, and 1058—459 = 599°, which is the corresponding 
thermometric temperature. But this temperature is greatly re- 
duced by the time the air reaches the motor cylinder, and still 
more heat is lost in the cyHnder before its work is done. To reheat 

* That is, the efiQciency of the compressor is assumed to be 50%. 



^60 co:mpress^ed air plant 

the air to a temperature which would double its volume in the 
motor cylinder would require reheating to a temperature much 
higher than 599^. If the temperature be raised by the reheater 
to 400" F. a loss of, say, 100° should be allowed for cooh'ng 
before the air is actually used. The absolute cylinder tempera- 
ture is then 300+459 = 759°, and the added volume of com- 
pressed air practically available is: 

529: 759 :: i: 1.43 + - 

That is, by reheating to 400° F., there has been an effective 
increase of about 43% in the volume of compressed air. For 
proper lubrication, a higher temperature would be undesirable 
in the motor cyUnder, and no material increase in economy 
could be realized in the operation of the motor. In practice 
the gain from reheating is considerably less than is here shown. 
For machine-drills and small, single-cylinder pumps, taking air 
at nearly full stroke, the increase of work ranges from, say, 
30-35^, without deducting the cost of the fuel used in the 
reheater. Expansive-working engines show a higher efficiency. 
For some purposes the determination of the quaatity of 
heat to be imparted in reheating is based on the temperature at 
which the air leaves the compressor cylinder, the idea being to 
recover the heat subsequently lost in cooling. Suppose, for 
example, that the compression is practically adiabatic, as is 
usual in single-stage compressors. Taking as the unit i lb. 
of air, or 13.2 cu.ft., at 65° F., and compressing to 70 lbs. gage, 
the heat of compression is: 

T' = T(|y"" = 65° +459°(^^™j°'" = 869° absolute, 

and the final thermometric temperature is, 869° — 459^ = 410° F. 
The rise in temperature due to compression is therefore: 

410 -65 =345 F. 

If the compressed air be subsequently cooled to 65°, its 

14. 7X13. 2 

volume becomes: = 2.29 cu.ft. In usmg this air 

54.7 



REHEATING COMPRESSED AIR 261 

without reheating and non-expansively, in a rock-drill having 
io% clearance, the work done is: 

W = (2.29X144X84.7X0.9) -(2.29X144X14.7) = 20,290 ft. lbs. 

But if the air be reheated to the final temperature of compression 
(345° F.), the work is: 

W' = — ^X 20,290 = 33,478 ft. lbs., and the gain by reheat- 
524 

ing is therefore: 33,478 — 20,290 = 13,188 ft. lbs., or 39%. 

The thermal cost of reheating this air will be: 345° X 0.23 7 5 
(specific heat of air at constant pressure) =81.9 thermal units 
(B. T. U.), equivalent to 81.9X778 = 63,718 ft. lbs. of work. 
Hence the efficiency of reheating in this case is : 

13,188-^63,718 = 20.7%. 

A working test, by Prof. Alex. B. W. Kennedy, on a reheater 
supplying air for a lo-H.P. motor, gave the following results: 
The air was reheated to 315° F., by burning about 0.39 lb. coke 
per indicated H.P.-hour, producing an increase of about 42% 
in the volume, and, if the efficiency had remained the same as 
during the trials with cold air, there should have been a decrease 
of air consumption in the ratio i -M .42 =0.7. The volume of cold 
air used (admission temperature, S^"^ F.) was 890 cu.ft. per 
H.P.-hour; the volume when reheated was 665 cu.ft., or 75%; 
so that the full economy from reheating was nearly realized. 

A summary of the results from two experiments on the 
above plant with cold, and two with reheated, air show:* 

1. With cold air. Incomplete expansion, wire-drawing, and 
other such causes, reduced the actual horse-power of the motor 
from 0.50 to 0.39. 

2. By heating the air to about 320° F. the horse-power at the 
motor was increased to 0.54. The ratio of gain due to reheating 
was therefore 0.54-^0.39 = 1.38. 

3. Deducting the value of 0.39 lb. coke per H.P.-hour, used 

* " Experiments upon the Transmission of Power by Compressed Air in Paris," 
Van Nostrand's Science Series, No. 106, p. 35, 



262 



COMPRESSED AIR PLANT 



in heating the air, the net efficiency becomes 0.47, instead of 
0.54. and the ratio of gain is reduced to: 

0.47 -^0-39 = 1-205. 

These experiments, though not made \\ith a well-designed 
reheater, show a substantial net gain from reheating. \Miere 
reheating is employed in mines, however, the quantity of heat 
imparted to the air is usually much less than that indicated 
above. Good results may be obtained by the application of 
even less than 100° F. 

The results of experiments by Riedler and Guttermuth. 
on the consumption of reheated air by an ordinary single-cylin- 
der 80 H.P. engine, are given in Table XX\TII.* This engine, 

T.\BLE x::s:\Tii 





Temperature Or Air. r^ c ,^ ^-^ 

Lonsumption 


Indicated 
Horse-Power. 




Test. 


Admission. 
Deg. F. 


riee .-vii yei 
Discharge. ^.R-Hr. in 
Deg. F. ^"•"• 


Efficiency. 


I 
2 

3 
4 


264. 2 
305 -6 
320.0 
338.0 


69.8 , 462.77 

84.4 1 431 09t 

95.0 ' 418.55 

120.2 432.12 


72.3 
72.3 
72.3 
65.0 


0.89 
0.90 
0.91 
0.87 



with CorHss valve gear, was originally designed and used as a 
steam engine, and no changes were made for adapting it to 
work with compressed air, except that the cyhnder was jacketed 
by the hot air on its way to the valve chest. The initial pressure 
was 95.5 lbs. absolute and the temperature of the air in the 
reheater did not exceed 338° F., at a coke consumption of 0.176 
lb. per H. P. -hour. * 

Construction and Operation of Reheaters. The reheater 
employed in the experiments referred to in the preceding para- 
graph consisted of a double cylindrical box of cast iron. 21 ins. 
diameter by 7,1, ins. high, enclosed in a sheet-iron casing. The 
air traversed the annular space between the cyHnders, being 
compelled by baffle-plates to circulate so as to come into con- 

* Wm. Cawthorne Unwin, ibid., p. 104. 



REHEATING COMPRESSED AIR 



263 



tact with the whole heating surface. The products of combus- 
tion from a coke fire in the inner cyhnder passed downward over 
the exterior surface of the annular air chamber on their way to 
the chimney. This served for a 10-12 H.P. motor. In an- 
other form of reheater the air passes through a coil of wrought- 
iron pipe, enclosed in a cylindrical casing. This gives a large 
heating surface, but wrought-iron pipe burns out rapidly unless 
the fire is kept moderate. Cast iron is preferable. 




Fig. 126. — IngersoU-Rand Reheater. 



The IngersoU-Rand reheater (Fig. 126) consists of two con- 
centric cast-iron shells, one within the other, the joints being 
packed with asbestos gaskets. The inner chamber forms the 
top of the fire-box. In shape this reheater is a truncated cone, 
set on a cylindrical fire-box, the cold-air main being connected 
by a flange coupling at the top and the hot air discharged near 
the base. Dimensions: 42 ins. outside diameter at the base by 



264 COJSIPRESSED AIR PLANT 

54 ins. high, with a grate 19 ins. diameter. It is stated that 340 
cu.ft. of free air per min., at 40 lbs. pressure, can be heated to 
360° F., with a gain of 30-35% in the energy developed. To 
reheat more than 400 cu.ft. of free air per min., 2 or more heaters 
are set in series, the air passing from one to another, allowing a 
maximum of 400 cu.ft. for each. 

The inner and outer shells of cast-iron reheaters are subjected 
to considerable differences of temperature, and the upper and 
lower ring joints between the shells are often difficult to keep 
tight. In the old Rand reheater (Fig. 127), the castings are 
more complicated in shape, the air passing between them in a 
thin sheet, from the inlet on the side to the discharge at the top 
of the dome. To provide for expansion and contraction, the 
joint above the fire-box is made by a stuffing ring and packing. 

The Sullivan reheater (Fig. 128) consists of a vertical coil 
of cast-iron piping, or hollow rings, encased in double sheet- 
steel shells, the space between the latter being filled with asbestos. 
Below is the grate and combustion chamber, the gases from which 
pass through the spaces between the air rings. To minimize 
leakage, the centers of the rings are joined by malleable-iron 
nipples, so that all expand and contract together. These heaters, 
usually designed for burning coal, coke, or wood, are made in 
3 sizes, for 345-800 cu.ft. of free air per min., having from 3-7 
rings, and measuring from 5 ft. 8 ins. to 7 ft. 6 ins. in height, 
by 33 ins. outside diameter. 

Internally fired reheaters, in which the air is heated by direct 
contact with the fire, are unsuccessful, because dust and injurious 
products of combustion are carried into the cylinder of the air 
motor. This trouble does not exist to the same extent when 
gasolene or other oils are used, instead of solid fuels, and not at 
all in the electric reheater, which, however, has rarely been used. 
Most reheaters have no provision for regulating the heat accord- 
ing to the variation in consumption of air, as in running machine 
drills, channellers, hoists, and other intermittently operating 
engines. This want of regulation is less important for constant 
running engines, like pumps. 

As the air chamber, in all externally heated or " dry " 



REHEATING COAIPRESSED AIR 



265 



reheaters, forms in reality a part of the air main, reheating 
increases the pressure only in a small degree. Its real effect 
is to increase the volume of air, which backs up in the main, 
reducing incidentally the velocity of flow and therefore the loss 
of pressure due to friction. The reheater should be placed as 
close as possible to the machine using the air. This is readily 




Fig. 127. — Old Rand Reheater. 



Fig. 128. — Sullivan Reheater. 



done with stationary engines, and even in the case of movable 
machines, like quarry channellers, the reheater may be set 
on the same carriage. 

With compound air-engines the exhaust from the high-pres- 
sure cylinder is sometimes passed through a reheater before 
going to the low-pressure cyHnder. A further benefit may be 



266 COMPRESSED AIR PLANT 

derived by injecting into the reheater a very small quantity of 
water. The specific heat of water is higher than the specific 
heat of air; also such part as is converted into steam gives up its 
latent heat in the air-engine cylinder and prevents trouble 
from freezing, even with a high rate of expansion. Similarly, 
it is beneficial to inject a little warm, or even cold, water into the 
feed-pipe of an air motor. Water thus used acts incidentally 
as a mechanical scourer, in removing accumulations of ice from 
the ports.^ 

It is evident from the construction of reheaters that the 
calorific power of the fuel burned in them is not economically 
utilized. The flue loss is large for the same reasons that apply 
to the work of ordinary shell or tubular boilers. But the thermo- 
dynamic advantage gained is so marked that the low efficiency 
of the reheater itself, in burning the small quantity of fuel 
required is of secondary importance. 

Reheaters for Underground Work. In the ordinary opera- 
tions of mining the reheating of compressed air has a limited 
application. For portable machines like rock-drills, continually 
being shifted from place to place, the use of reheaters is economic- 
ally out of the question, because they would have to be moved 
about with the drills, to prevent the reheated air from losing its 
heat and temporary increase of volume. 

For stationary engines, however, as underground pumps, 
hoists and rope-haulage engines, where the reheater can be 
placed permanently close to the engine, reheating may be 
applied with a decided gain in efficiency. Incidentally it pre- 
vents the accumulation of ice in the exhaust ports. It may be 
difficult underground to arrange for burning fuel under a reheater, 
notwithstanding the small quantity required, because of the 
vitiation of the mine atmosphere. Where the conditions are 
such that strong combustion is not allowable, it will still be 
found that some advantage is obtainable from air engines by a 
very slight added temperature — say, only 25°-5o° F. The use 
of the internal electric reheater, already referred to, in which a 
resistance coil is placed in an enlarged section of the air main, 
avoids the difficulty of disposing of the products of combustion 



REHEATING COMPRESSED AIR 267 

of fuel and would be especially useful in gassy collieries. Another 
mode of applying electric reheating is to wrap the resistance 
coils around a short length of the air pipe. 

At the North Star Mine, Grass Valley, Cal., a reheater has 
been placed on the surface near the shaft mouth and the com- 
pressed air carried underground by a pipe covered with non- 
conducting material. While some saving can thus be realized 
for moderate distances, it is not practicable for long pipe lines. 
In any case, non-conducting coverpg should be used for the pipe 
from reheater to air engine, however short the distance. In a 
case on record,* where this distance was only 20 ft., without 
pipe covering, the gain in power was only 12%, though the 
absolute temperature of the air at the reheater was increased 
38%, with corresponding theoretical increase of volume. For 
operating an underground pump in another California mine, 
the air is reheated by steam conveyed from the surface. Steam 
may thus be used to greater advantage than if employed directly 
in the cylinder of a pump; for, in condensing, the latent heat 
raises the temperature of the air and is so converted into work. 

Reheating is important in connection with the operation 
of surface or underground hoisting engines by compressed air 
(see latter part of Chap. XVII). 

Following are some results of experiments by Prof. J. T. 
Nicholson, in reheating air from the Taylor Hydraulic Air 
Compressor, at Magog (Chap. XV). The air was used in a 
27-H.P. Corliss engine, at a pressure of 53 lbs. There were 5 
tests: (i) with cold air; (2) reheating by steam injected into 
the air main near the engine; (3) the compressed air was passed 
into a steam boiler, and heated by mixing with the water and 
steam; (4) the compressed air was blown upon the surface 
of the water in the boiler, and heated by mixing with the steam; 
(5) the air was passed through a tubular reheater, fired by coke. 

Without reheating, 850 cu.ft. of free air were used per I.H.P.- 
hour. By reheating in the boiler, a mixture of 10-15 lbs. of 
steam with the air reduced the consumption of air from 850 
cu.ft. to from 300-500 cu.ft. per I.H.P.-hour. Thus, i added 

* Richards, American Machinist, Feb. 28, 1895. 



268 COMPRESSED AIR PL.\XT 

H.P. was obtained by wet heating, at an expenditure of 1-1.3 Ibs. 
o£ coal per H.P.-hour. By dry heatimg in the coke-fired reheato", 
the air was raised to 28 7" F. At this temperature, 640 cu.ft- 
of free air were required per H.P.-hour. or 210 cu.ft- less than 
with cold air, the saxing in quantity- of air being about 2^~^. 
By burning in the rehe^rer 47 crs. coke per hour. 100 H.P. in 
cold compressed air was raised to 133 H.P.. making an expendi- 
ture of 1^2 lbs. coke for each added H.P.-hour. Though these 
results indicate :hi: the reheater used was not ver\' eflScient, 
the fuel consumption is far lower than is attainable in the best 
boiler and enr ictice. 

In a paper u\ Liarence R. We\"mouth. on ** Reheating Com- 
pressed Air with Steam." a detailed discussion is given of the 
thermodxTiamics of this mode of procedure, with deductions as 
to its efficienc\\ The author gives the results of injecting steam 
into the air main, and of p>assing the compressed air into a 
boiler. 



CHAPTER XX 

COMPRESSED AIR ROCK-DRILLS * 

This chapter will be devoted to the description of a number 
of representative drills, together with notes on the performance, 
air consumption and operation of machine drills. For any 
rock harder than soft shales, coal, and other similar material, 
the percussion drill only is of practical use. All attempts to 
construct a rotary pneumatic rock-drill have thus far failed. 
The diamond, and other core, drills for deep boring, and the 
Brandt rotary drill, operated by hydraulic power, have no place 
in the present discussion. 

General Description. The reciprocating or percussion rock- 
drill, except those machines that operate on the hammer prin- 
ciple (Chap. XXI), consist of a cylinder, in which compressed 
air or steam is used, the drill bit being clamped to a chuck on 
the end of the piston rod (Fig. 129). Many different valve 
motions are in use. Some resemble the valve motions of cer- 
tain single-cylinder pumps; in others the valve is positively 
moved by a tappet, actuated by the strokes of the piston. 
The necessary rotation of the drill bit on its ax's, for keeping the 
hole round and preventing the bit from sticking {" fitchering "), 
is produced automatically by a rifle-bar, ratchet and pawls. 
Working speeds are from 300-400 strokes per min. for the larger 
sizes of drills, to 500 strokes for the smaller; normal length of 
stroke, in drills of average size, 4J-6 ins. The admission of 
air, and therefore the speed and force of the blow, are controlled 
by a hand valve in the air pipe close to its connection with tlie 
valve chest. 

A feed screw, with crank and handle, is carried in a bearing 

* Except as regards drill mountings, this Chapter deals with Reciprocating 
Drills (see General Classification, p. 273). For Hammer Drills, see Chap. XXI. 

269 



270 COMPRESSED AIR PLAXT 

at the rear end of the shell supporting the cylinder, and engages 
with a nut on the under side of the cylinder casting. The entire 
drill head is thus fed forward by hand as the hole is deepened. 
(An automatic feed, used to some extent for surface work, is 
neither necessary nor satisfactory for underground ser\4ce.) 
When the cyhnder has been fed forward as far as the length of the 
screw and of the supporting shell Tsill permit, the drill is stopped. 
By reversing the feed the cylinder is moved back on the shell, 
the bit is removed and a longer one clamped in the chuck. Ths 
cylinder is then fed forward until the new bit touches the bottom 
of the hole, \vith the piston nearly at mid-stroke; the air is turned 
on slowly and the work proceeds. The length of stroke, and 
therefore the force of the blow, are under the control of the 
drill-runner. If the drill is fed forward faster than the hole is 
being deepened the stroke necessarily shortens, because the bit 
strikes the bottom of the hole before the full length of stroke 
is reached; conversely, with too slow a feed, the piston will 
strike the front cyhnder head. Thus, the force of the blow^ 
may be regulated to suit the conditions. WTien starting a 
hole, the stroke should be short until the bit has adjusted itself 
to the shape of the bottom of the hole. For hard rock, a short, 
rapid stroke is best; a longer stroke for softer or tough rock. 

Drill Mountings. The drill head, comprising the cylinder 
and its appurtenances and the supporting shell, is mounted on a 
tripod or column. For surface w^ork the tripod only can be used; 
underground, either the tripod or column, depending on the 
size and shape of the working in which the driUing is to be done. 

I. Tripod. (Fig. 129). The legs, which are telescopic, are 
hinged by a hea\y bolt to the tripod head, and can thus be set 
as necessary for adjusting the position of the axis of the cylinder 
and bit, for the hole to be drilled. To the tripod head is bolted 
the '' shell," which has guides for supporting the cylinder, 
as it is fed forw^ard. After the machine is in position for drilling 
all bolts are tightened. Weights are usually slipped on the 
tripod legs, to prevent the drill from shifting while in operation. 

Tripod mountings are required for underground \vork, when 
the distance between roof and floor, or between the walls of the 



COMPRESSED AIR ROCK-DRILLS 



271 



workings, is too great to permit the use of columns. \\(here a 
choice exists, the tripod is sometimes preferred to the column, 




Fig. 129. — Tripod Mounting (Sullivan Drills). 

because it can usually be set up in less time and allows greater 
freedom in locating the holes, so as to produce the best results 



272 



COMPRESSED AIR PLANT 



under existing conditions as to character of rock and shape of the 
face. This may be true, also, in sinking shafts of large cross- 
sectional area, or when the rock is so irregularly fissured that a 
symmetrical arrangement of the round of holes is undesirable. 
2. Column or Bar (Fig. 130) is a heavy steel tube, 3-5 J ins. 




Fig. 130. — Double-Screw Column Mounting. 



diameter, according to the size of drill. It is set between the 
roof and floor, or the wails, of the working. The lower end of the 
column shown has a cross-piece or base, through which pass a 
pair of jack-screws. The upper end has a serrated head, which, 
by tightening up the screws, takes a fiirm hold upon the surface 



COMPRESSED AIR ROCK-DRILLS 273 

against which it is pressed; blocking and wedging are generally 
required. Another form of column, with a single, telescopic 
jack-screw, is useful in small tunnels or mine workings, and for 
shaft-sinking. In large workings the double-screw column is 
preferable. The drill is carried on an arm, attached to a 'collar 
which is clamped fast in any desired position, for adjusting the 
height and angular position of the drill. With the single-screw 
column the drill is attached directly to the collar,- the arm being 
omitted. Sometimes two drills are mounted on the same column. 

The column mounting is best for driving mine tunnels, 
cross-cuts, and drifts, and for advance headings of railroad 
tunnels. It is often employed, also, for s toping in veins less 
than about 7 ft. thick, when the dip of the vein and the 
method of mining make it difficult to use tripods. When set 
horizontally, as in shaft-sinking, the column is called a " bar." 
Shaft-bars of extra length, for wide shafts, have a pair of 
adjustable legs hinged at the middle point to carry the weight of 
the drill or drills without making the bar inconveniently heavy. 

Formerly, for driving railroad tunnel headings, 4, 6 or more, 
drills were mounted on a carriage on a track, but they are now 
rarely used. 

General Classification of Rock-Drills. (A) Reciprocating 
drills, comprising those in which the drill bit is firmdy attached 
to the piston rod; (B) Hammer drills, in which the bit does not 
reciprocate, but is held in a socket in the forward end of the 
cylinder, and is struck by the piston as by a hammer.* 

Reciprocating Drills 

• Classification based on the design of the valve-motion if 

A. Air- thrown valve machines (nearly all are of the spool- 
valve type). 

B. Tappet-valve machines. 

C. Miscellaneous types. 

Examples of each form are given below. 

* Air Hammer drills are discussed in Chap. XXI. 

t For another classification, see Eng. df Min. Jour., Feb. 22, 1913, p. 421, 



274 COMPRESSED AIR PLANT 



AIR-THROWN VALVE MACHINES 

*• Sergeant" Drill. Fig. 131 is a longitudinal section. The 
spool-valve and main air and exhaust ports are somewhat 
similar to those of a single-cylinder pump. Air is admitted 
on one side of the valve chest, the exhaust opening being on the 
other side. 

The valve-motion consists of two parts: a spool- valve, 
which controls the main cyHnder ports, and an arc-shaped 
tappet, set in a curved slot or seat, between the cylinder and 
valve chest. The ends of this tappet, projecting sHghtly into 
the main cylinder, are struck alternately by the front and back 
shoulders of the large annular recess in the middle of the piston, 
thus causing the tappet to oscillate at each stroke of the piston. 
Alongside of the tappet, and in its seat, are three auxiliary ports, 
one in the middle and one near each end of the seat. These 
ports connect with the spool- valve chest above; the middle port 
with the middle of the chest, the rear port {i.e., nearest the back 
end of the cylinder) with the forward end of the chest, and the 
forward port with the rear end of the chest. In the face of the 
tappet is a curved slot, just long enough, when in the extreme 
positions of its throw, to form a communication between the 
middle auxiliary port and one of the end auxiliary ports. There- 
fore, at each stroke of the piston, the tappet places the opposite 
end of the valve chest in communication with the exhaust, 
thus causing the throw of the spool-valve. In the peripheral 
surface of the spool-valve is a very fine longitudinal slot, which 
constantly admits a small quantity of live air to both ends of 
the chest. Hence, when either end of the chest is connected 
with the exhaust, the valve is thrown towards that end by the 
air pressure in the other end of the chest. 

In Fig. 131 the piston is beginning its forward stroke. The 
spool-valve, in its rear position, is admitting air to the back end 
of the cyHnder, while the forward end is connected with the 
exhaust. As the piston advances, the rear shoulder of the 
annular recess in the piston strikes the projecting end of the 
tappet and throws it over. This changes the relation between 



COMPRESSED AIR ROCK-DRILLS 



275 





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C3 



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O 



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;/3 



6 



276 



COMPRESSED AIR PLANT 



the auxiliary ports, described above, exhausts the air from the 
front end of the chest and throws the spool-valve forward, thus 
preparing for the back stroke of the. piston. 

The use of the arc-tappet avoids in large part one of the 
defects of the ordinary spool-valve drills, viz., irregularities 
in the operation of the drill, caused by wxar of the valve and seat, 
which permits leakage of air past the valve. Adjustments for 
irregularities of stroke produced by wear are made by the com- 
pensating device shown in Fig. 132, which is an enlarged section 
of the spool- valve and chest. A hollow brass plug P, having a 
very small hole H, permits the passage of a little live air to the 
back end of the chest. Should the piston strike the back cylinder 



Front Encf. 




Fig. 132. — Spool- Valve and Chest. " Sergeant " Rock Drill. 



head, the area of H is reduced slightly by peening (hammering) 
up the edge of the hole. This decreases the quantity of air 
passing to the end of the chest, throws the valve a little later, 
and so increases the cushioning in the rear end of the cylinder. 
If the stroke be too short, U may be found partly obstructed and 
should be cleaned, to admit more air to the end of the chest; if 
the stroke be still too short, the area of B. is slightly increased 
by reaming out with the point of a knife blade. 

Rotation of piston and bit: A rifled bar, with a ratchet head 
and pawls set in the rear cylinder head, engages with a corre- 
spondingly rifled nut screwed into the end of the hollow piston. 
The ratchet teeth and pawls (Fig. 133) are so shaped that, on 
the forward stroke, the piston moves without rotation, the 



COMPRESSED AIR ROCK-DRILLS 



277 



rifle-bar turning the ratchet 
in its seat. On the back 
stroke the pawls prevent the 
ratchet from turning, so that 
the piston is compelled by 
the rifle-bar and nut to rotate 
through a part of a revolu- 
tion. The ratchet ring, with 
internal teeth, with which 
the pawls engage, is not 
fastened rigidly in the back 
cyhnder head, but is held by 




Fig. 133. — " Sergeant " Ratchet and 
Rifle-Bar. 



friction only, under pressure 
of a cushion spring in the 
back head of the cylinder. 
Hence, when the drill bit 
sticks in the hole, or for any 
reason cannot rotate freely on 
the back stroke, the ratchet 
itself turns, thus preventing 
injury. The drill is fed for- 
ward on its supporting shell 
by a long feed screw, engaging 
with a feed nut in a lug on 
the under side of the cylinder. 
The Sergeant drill is built 
in seven sizes, the cylinders 
measuring: 2, 2|, 2§, 2f , 3, 3! 




o 
U 



u 



a 

C/2 



< 



■4-) 

Q 






CO 



o 

I— I 



278 COMPRESSED MR PLANT 

and 3I ins. diameter; weights of drillhead, unmounted, iio- 
405 lbs. 

Sullivan '' Liteweight " Drill (Fig. 134) has a three-ringed 
spool valve, thrown by air admitted from and exhausted directly 
into the cylinder, through auxiliary or reverse ports. These 
ports (not fully shown in the cut) are grooved in the valve-chest, 
instead of being cored in the cylinder casting, as in the older 
pattern of the drill. The middle ring of the valve is of larger 
diameter than the others. As the piston makes its stroke, the 
reverse ports put the ends of the chest into communication with 
opposite ends of the cylinder, thus throwing the valve and 
preparing for the next stroke of the piston. The valve chest is 
bushed with steel. The valve will still reverse even when the 



Pawl Spring 

Fig. 135. — Rotation Device, Sullivan *' Liteweight " Drill. 

drill head is fed down so that the piston stroke is shortened 
to only about ^ in. A steel liner, ^ in. thick, is pressed into 
the cylinder and dowelled to prevent movement. It gives a 
hard, smooth surface to take the piston wear, and enables a 
lighter cylinder casting to be used. An automatic lubricator 
is provided at the rear end of the cylinder. 

The ratchet of the rotation device (Fig. 135) has rounded 
toothed surfaces, engaging with two small steel rollers, instead 
of angular pawls. These rollers are held in place by peripheral 
springs, as shown. The ratchet ring and collar are held against 
the rear cyhnder head by friction only, to permit slippage in 
case the bit wedges in the drill hole. 

This drill is made both with and without a water injection 
attachment, for keeping the bottom of the hole clean by driving 
out the sludge and thus increasing the driUing effect. Fig. 136 



COMPRESSED AIR ROCK-DRILLS 



279 



shows the water t>i)e of drill.* The piston, piston rod and 
rifle-bar are bored out to permit insertion of a small tube, 
through which water and compressed air pass under pressure 
into the hollow bit to the bottom of the hole. The water supply 
is from an i8-gal. tank, connected by hose to the compressed- 
air pipe and to the drill. A gravity supply of water may re- 
place the tank, provided the water pressure does not exceed 
the air pressure. Incidentally, the use of the water attach- 
ment saves the time ordinarily required to clean out the drill 
hole at intervals. 



Throttle Valve 




Fig. 136. — Sullivan Drill Mounted, with Water Tank and Connections. 



The " Liteweight " drill is made in 3 sizes: 2f , 3! and 3f-in. 
diameter of cy Under. The 2f-in. size is a one-man machine, 
weighing unmounted, 162 lbs. 

Wood Drill (Fig. 137) is another example of spool-valve 
machine. With a few exceptions, its main features are similar 
to those of the Sulhvan drill, already described. The front and 
back vertical reverse ports for the valve are bushed, as shown, 
and, by means of horizontal channels cut in the lower part of 
the valve-chest casting, communicate with opposite ends of 
the chest. By the reciprocations of the piston the ends of the 

* It should be stated that this device has been used for some years in the 
" Water-Leyner," now the " Leyner-Ingersoll," drill (Chap. XXI). 



280 



COMPRESSED AIR PL.\XT 




c 



chest are thus connected 
with the cyHnder, and 
thence through the main 
ports back to the chest 
and finally to the exhaust 
opening of the machine. 
The rotation of the piston 
and bit are positive, the 
ratchet ha\ing no slip 
ring. The four pawls are 
set in a solid forging 
(pawl-holder) , which is 
held rigidly by dowel 
pins, as sho^^^l, between 
the pawl-holder and rear 
cylinder head. 

This drill is made in 
6 sizes for underground 



2, 2 



3i-in. 



diameter of 



^ ser\ice 
f and 

cylinder; weights, un- 
I mounted, 85-390 lbs. The 
^ two smallest sizes are 
I one-man machines. 
5 Climax Imperial Drill 

o (Fig. 138). This is \ 
well-known English ma- 
chine. Air enters the 
valve-chest by the air tap, 
and thence passes into the 
annular recess b of the 
valve, which, by its re- 
ciprocations, opens com- 
munication through the 
valve-seat ports c, alter- 
nately ^^^th the main 
cyHnder ports a, a. In 



COMPRESSED AIR ROCK-DRILLS 



281 




282 COMPRESSED AIR PL.\XT 

the valve are the recesses J. sho\\Ti also in section to the left. 
These, by the movements of the valve, control the exhaust 
ports e, which connect with the main exhaust on the side of the 
chest. The valve is thro\\Ti by admitting a Uttle Uve air through 
the small grooves jj\ to the ends of the chest, this air being 
alternately discharged by the larger auxiliar}' ports /./''. These 
open into the cylinder through the auxihary ports g. exhausting at 
each stroke into the annular recess of the piston and thence into 
one of the square ports //. leading to the main exhaust. Ports 
g are bushed with composition metal rings, shaped at the lower 
end to fit closely upon the piston. As shown in the cut. the 
piston is in position to begin its forward stroke; the valve 
has been thro'^Mi to the right and is admitting air to the rear 
end of the cylinder. The drills are designed to nm at the high 
speed of from 450-500 strokes per min. 

A specialty of this drill is the " dust allay er,'' which is 
attached to the air tap by a nipple and cup. forming a ball-and- 
socket joint. It is. in efi'ect. an ejector, drawing water from any 
convenient source, by means of a small quantity of compressed 
air led from the throttle. By the same air the water is sprayed 
forward into the mouth of the drill hole. 

The " CHmax " drill is made in 7 sizes: if, 2, 2I, 2|, 3, 3J 
and 3i-in. diameter of cyhnder. 

Holman Drill, an Enghsh machine, is made in 2i, 2J, 2^, 
^h 3i' 3i ^nd 3f-in. sizes. In a few of their details the smaller 
sizes differ somewhat from the larger, though the general design 
is the same in all. Fig. 139 illustrates the sizes from 2J to 2|-in. 
The movements of the spool-valve i, which control main air 
and exhaust ports of the usual form, are caused as follows: 
Below each end of the valve-chest, and communicating from 
chest to cylinder, is a short vertical port, with a coned or taper 
seating. In each of these ports is a pair of steel balls 4, 5. the 
former of which controls the auxiliary port 7. Both balls are 
under the pressure of the spiral spring 8. The seat is so shaped 
that the lower and smaller ball 5 vdW project slightly into the 
cyhnder. whenever permitted to do so by the position of the 
annular recess 6. around the middle of the piston. Hence. 



COMPRESSED AIR ROCK-DRILLS 



283 



by each piston stroke 
the lower ball 5 receives 
a slight upward blow 
from the inclined 
shoulder of the recess. 
This lifts the larger ball 
4, thereby opening aux- 
iliary port 7, and plac- 
ing the corresponding 
end of the valve-chest 
in communication with 
the main exhaust 3. 
Owing to the pressure 
of the air occupying 
the opposite end of the 
chest, the spool- valve is 
then reversed, to prepare 
for the next stroke of 
the piston. Ball valves 
are not Hable to break- 
age, and, as they receive 
a slight rotary motion 
from each blow of the 
piston, the wear tends 
to be equalized, thus 
keeping them round and 
preventing leakage be- 
tween the balls and their 
seats. 

I n g e r s 1 1 - R a n d 
** Butterfly" Drill (Figs. 
140-143). The valve 
(Fig. 141) is " air- 
thrown," having no 
mechanical connection 
with the piston. It con- 
sists of two fiat wings, 




284 



COMPRESSED AIR PLANT 



with ground seating surfaces, these wings forming one piece with a 
central trunnion. The valve oscillates slightly in a groove 
or slot in the chest, being actuated by the unbalancing of air 
pressure alternately on the wings. There are 4 ports, opening 
into the faces of the valve slot, a separate inlet and exhaust 




Fig. 140. — Ingersoll-Rand " Butterfly " Drill, for Tripod or Column Mounting. 

for each end of the cylinder. The two inlet ports open opposite 
to each other at one end of the slot, the exhaust ports being 
similarly placed at the other end. When one of the valve 
wings closes the inlet to one end of the cylinder, the opposite 
face of the other wing closes the exhaust port from the other 



r 







Fig. 141.— Valve of "Butterfly" Drill. 



end of the cylinder. One inlet and one exhaust port are there- 
fore always open. 

Fig. 142 shows diagrammatically the operative conditions 
on beginning the forward stroke. The rear inlet port ^2' and 
the forward exhaust port Ei are open, the others being closed. 



COMPRESSED AIR ROCK-DRILLS 



2B5 



Since the piston in this position covers the rear exhaust port £2, 
the pressure of the inlet air holds the valve to its seat over port Si. 
When the advancing piston uncovers E2, live air passes through 
the cylinder to £2, almost balancing the pressure on the two 
valve wings. But, the resultant of the forces acting on the 
valve, comprising the flow of the exhaust air in £1, the impact 
of the inlet air on entering the chest, and the friction of the air 
in passing through 52 and £2, keeps the valve on its seat, in its 
first position, until the forward stroke is nearly completed. 
When the piston passes over and closes the forward exhaust 
port £1, the cushion pressure produced in front of the piston 
acts through the port Si, and throws the valve to the position 
shown by Fig. 143. The back stroke now begins, live air enter- 



Inlet 



Inlet - 




Fig. 142. 



Fig. 143. 



ing through Si, while the exhaust escapes through £2. The 
cycle of operation is then completed, as already described. 

The '^ Butterfly " drill is a high-speed machine, making 500 
-600 strokes per min. The ports are large and with a small 
amount of movement the valve gives a large port area. Tests 
on the earlier drills of this type indicated that the air consumption 
is rather high. A new 2|-in. drill showed a consumption of 
95 cu.ft. free air per min. at 70 lbs. pressure and 125 cu.ft. at 
,85 lbs. This is due to the fact that, when the valve reverses, 
there is momentarily a direct communication from inlet to 
exhaust, as from S2 to £2. As the valve trunnions and faces 
wear, the seating may become imperfect, so that notwithstand- 
ing the quick action of the valve there is a loss of air. For 
example, a new 3-in. drill, tested at 70 lbs. pressure (with an 
Excelsior air meter), used 120 cu.ft. free air per min. The 



286 



COMPRESSED AIR PLANT 



same drill, with valves and valve-chests 3 months and 5 months 
old, consumed respectively 145 and 170 cu.ft. Later improve- 
ments have largely overcome these difficulties and lessened the 
air consumption. 

This machine is a fast driller and marks a distinct advance 

in simplicity of construction; it 
" muds " well, and, being a one- 
man drill, the labor cost is low. 
The " Butterfly " valve is used 
also in the Leyner-Ingersoll and 
Ingersoll-Rand hammer drills 
(Chap. XXI). 

*' Chicago Gatling" Drill 
(Chicago Pneumatic Tool Co.) 
affords another example of an air- 
thrown valve, not of the spool t}^pe 
(Fig. 144). The valve is a 2-oz. 
hollow steel ball, held in a cyHndri- 
cal pocket or cage, taking the 
place of the usual chest. Around 
the periphery of the cage are 12 
small apertures for admitting air. 
The ball, with a J-in. throw, con- 
trols the main ports to the cyhn- 
der; seating surfaces are of steel. 
Live air pressure holds the ball 
on its seat until a drop in pressure 
on the other side of the ball is 
caused by the piston uncovering 
the corresponding exhaust port. 
That is, the valve movement is 
not dependent upon compression; 
the ball is thrown by the differ- 
ence in total pressure on the oppo- 
site sides, the exhaust opening being larger than the apertures in 
the valve cage, through which live air enters. A similar valve 
motion is used in the Chicago ^' Hummer " drill (Chap. XXI). 




o 
U 

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o 
H 



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o 

fcXJ 

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a 
O 
o 

a 

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o 



COMPRESSED AIR ROCK-DRILLS 



287 



Speed of stroke, 600-700 per min. ; the stroke is uncushioned. 
Size of drill, 2f -in. ; weight, unmounted, 200 lbs. 

Other Spool-Valve Drills are built by the McKiernan- 
Terry Drill Co., N. Y., the Chicago Pneumatic Tool Co., and 
the Cochise Machine Co., Los Angeles, Cal. The '^ Chicago 
Slogger " has an arc-shaped auxihary tappet valve, resembhng 
that of the " Sergeant " drill (Fig. 131) . 

The Siskol drill was one of the four successful competitors 
in the fourth (1909) Transvaal stope-drih contest, which was 
conducted underground, under regular working conditions, 
and lasted nearly a year. It is built in two sizes, 23^-in. and 
2^-in. diameter of cylinder. Fig. 145 shows the one-man 




Fig. 145. — Siskol Drill (2j^-in. diam. of cylinder). 



23^-in. stoper, which weighs, unmounted, about 120 lbs. 
During the test run, with two of these machines, 14,083 linear 
ft. of hole were drilled in 215 8-hour shifts; average speed, 
0.818 in. per min. per drill. The other successful competitors, 
Holman 2|-in. and 2f-in. (Fig. 139), and Chersen (see below), 
also working in pairs, drilled respectively 12,779, 11,744 and 
11,781 ft. of hole in the same time, average speeds, 0.742, 
0.682 and 0.684 in. per min. per machine. In another test of 
the 2j^-in. drills, in May, 191 1, lasting ij hours, 14.06 cu.ft. 
of free air were used per Knear in. of hole; air pressure, 75 lbs. 
The Chersen drill is a 2f-in. stoper, which also gave good 
results in several Transvaal contests. It is peculiar in having 
the spool valve chest placed transversely across the cylinder, 



288 



COI^IPRESSED AIR PLANT 



It 



k' 



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/ 



r^ 



1 
1 



\ 



^ 



a 

d 



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c/: 



y 



The vaive has a very short 
stroke — 15 one-thousandths of 
an inch — and is actuated by a 
sHght cushioning of the air in 
the cyHnder, at each end of the 
stroke. Owing to its small 
travel, the action of the valve is 
apt to be interfered with by the 
entrance of grit into the chest. 
This difhculty was clearly 
brought out in the Transvaal 
tests. 



TAPPET-VALVE MACHINES 

Sullivan '*Hy-Speed" Drill 

(Sullivan Machinery Co.) has 
a spool-valve which controls the 
main cyhnder ports, and a small 
auxiliary slide valve, thrown by 
a 3-arm tappet or rocker (Fig. 
146). The tappet oscillates in 
an arc-shaped slot, as its lower 
arms are struck alternately by 
the beveled shoulders of the 
piston. On the back of the 
tappet is an arm of standard 
rack-tooth form, which engages 
with a socket in the underside 
of the slide valve, thus throwing 
this valve positively. The slide 
\'alve controls small auxiliary 
ports, communicating with the 
spool- valve chest, thus causing 
the spool to reciprocate. An 
advantage of the flat valve is 
that it is held to its seat by air 



COMPRESSED AIR ROCK-DRILLS 



289 



pressure, SO that wear does not cause leakage, 
tappet here used is an improve- 
ment on the design used for 
many years for all tappet drills; 
breakage of the tappet is greatly 
reduced. Automatic lubrication 
is provided, as in the " Lite- 
weight " drill. 

The rotation, feed mechan- 
ism, and water attachment, are 
the same as in the Sullivan 
'' Liteweight '^ drill (Fig. 135). 
Drill sizes, 2|, 3, si, si ^^nd 4i- 
in. diameter of cylinder ; weights, 
unmounted, 250-730 lbs. 

The 3! -in. size is made in two 
styles. The heavier (weighing 
730 lbs.) is the " Engine-feed " 
drill, designed for drilling 3-in. 
holes to 20 ft. deep in quarry- 
ing and other open-cut work. 
Its valve-motion and rotation 
are as described above, but it is 
mounted on a special shell, to 
give a continuous feed of 4! ft. 
Due to its unusual weight, a 
small 2-cylinder hoisting engine 
is attached to the back of the 
shell for feeding and raising the 
drill. This engine is geared to 
the top of the feed screw, which 
also has a handle, for feeding by 
hand, if desired. 

Murphy ** Little Champion" 
Drill (Fig. 147). The rotation 
mechanism is similar to that of 
other drills already described. 



The non-pivoted 




290 COMPRESSED AIR PL.\NT 

The upper arm of the 3 -arm tappet a engages with the flat sHde 
valve h. This valve controls the main ports ex and the exhaust 
port, which is opposite the middle of the tappet. A novel 
feature of the drill is the mode of producing the tappet's move- 
ments. Its two lower arms do not project into the top of 
the cyhnder bore, to a contact with the piston, as is the case with 
most drills of this t}^)^ Instead, a steel pin d, sHding in a bushed 
opening in the cyhnder casting, is set under each tappet arm. 
As the piston makes its strokes, the curved shoulders of the annu- 
lar groove e, in the piston, alternately strike the rounded ends 
of the forward and rear tappet pins, d,d, thus pushing up the 
pins and causing the tappet to oscillate. 

There are eight sizes: 2 J, 2\, 2f, 3, 3^, 3^, 3I and 3|-in. 
diameter of cyhnder, the drill head and shell weighing, un- 
mounted. 125-395 lbs. 

Holman Drill is an Enghsh tappet- valve machine, made in 
Camborne, Cornwall (see also Hohnan spool- valve drill). A 
3-arm pivoted tappet, oscillated by shoulders on the piston, 
throws a D slide valve, which is held on its seat by a plate 
spring bearing against the valve-chest cover. 

MISCELLANEOUS TYPES 

Temple-IngersoU *' Electric-Air " Drill (Fig. 148) is unique 
in the mode of combining both systems of powxr transmission 
and in no way belongs to the class of electric-driven drills, which 
for years have been brought out from time to time, but which 
as yet have not given wholly satisfactory results.* 

This machine comprises three parts: a drill, and an air 
pulsator driven by an electric motor. Pulsator and motor are 
mounted on a small truck, close to the drill and connected with it 
by two short lengths of hose. The drill differs materially from 
an ordinary rock-drill; it is valveless, and the cyhnder is of 
larger diameter; the short piston, with packing rings, resembles 
the piston of a steam engine. The drill may be mounted on a 
column or tripod, and is pro\ided with a feed screw and rotation 

* Some resemblance to this machine is traceable in the design of the " Pneum^ 
electric Coal Puncher " (Chap. XXII), 



COMPRESSED AIR ROCK-DRILLS 



291 



device. It has no buffers, springs or side rods. The pulsator 
is in effect a duplex, single-acting compressor, with cranks at 




w 



to 

c 



fa 

H 



00 



1 80°, the crank-shaft being driven through single-reduction 
gearing from the motor armature. From the pulsator one hose 



292 



COMPRESSED AIR PLANT 



passes to the back end of the drill cylinder, the other to the 
forward end. These connections serve as ports for admission 
and return of the compressed air. There is no exhaust; the 
air circuit is closed, the same air being used over and over. 
Thus the speed of stroke depends on the motor speed, which is 
varied by a controller, operated through a cord by the drill 
runner. If a direct-current motor be used, it is designed for 
three speeds; or an alternating-current, single-speed motor may 
be employed if desired. 

The pulsator runs at a low air pressure, only a small degree 
of compression being necessary for transmitting the power, the 
air acting as a spring between pulsator and drill. Incidentally, 
the air cushions the drill piston at the ends of the stroke. 
Leakage from joints and past the pulsator pistons is made up 
by a compensating valve (not shown in the cut), adjusted to 
maintain a practically constant pressure in the air circuit. When 
the pressure falls below the limit, the valve opens automatically 
and admits more air, which is compressed by the differential 
area between the two parts of the piston in the first cylinder, 
until normal pressure is restored. The pulsator cranks and 
pistons are lubricated by the " splash " method, the lower part 
of the crank-case being partly £lled with oil. Some oil is 
atomized and carried with the air into the drill cylinder. 

This drill is made in 4 sizes (Table XXIX), which, in working 
capacity, correspond approximately to the 2, 2f, 3 and 3i-in. 
sizes of the " Sergeant" drill (Fig. 131). The voltage recom- 
mended is 220. For alternating current the standard motors 
(which are stronger and simpler than those for direct current) 



Table XXIX 



No. 


Diameter 

Cylinder, 

ins. 


Stroke, 
Ins. 


Strokes 


Weight of 
Drill un- 


Weight of Pulsator 
and Motor, Lbs. 


Approx. 
H.P. at 


per Min. ' mounted, 
Lbs. 


D.C. 


A.C. 


Pulsator 
for I Drill. 


3-C 
4-D 
4-E 
5-C 


3l 
4l 
4! 

5i 


62 

7 

7 
8 


475 
415 
440 
400 


119 
223 
228 
299 


5^5 

645 

883 

1050 


370 

545 
928 
820 


3 

} ^ 

5^ 



COMPRESSED AIR ROCK-DRILLS 293 

are three-phase, 25, 30, 50, and 60 cycle. Direct-current 
motors, wound for 440 or 500 volts, may be used, but these 
voltages are unnecessary and somewhat dangerous for under- 
ground service. 

" Triumph " Drill (Fig. 149) is valveless. Air is admitted 
by the two-way throttle c, on top of the cylinder, entering thence 
the annular port d. On beginning the stroke, as in the cut, d 
is m communication with an annular recess e, near the forward 
end of the piston; whence the compressed air passes through/, 
which is one of four longitudinal ports in the body of the piston, 
to the rear end of the cylinder. The forward stroke then takes 
place, the air in front of the piston being exhausted through 
ports j. As the piston advances, exhaust ports j are covered 
by the solid part of the piston, thus cushioning the end of the 
stroke. When the annular recess e, in the piston, comes opposite 
ports y, the air exhausts from the back end of the cylinder. At 
the same time, the annular recess Ji, near the rear end of the 
piston, comes into connection with the inlet d, thus admitting 
live air through the four longitudinal piston ports, one of which, 
ij is shown. These ports conduct the air to the forward end 
of the cylinder and the stroke is reversed. The parts are so 
proportioned that the air acts at full pressure throughout only 
about one- third of the forward stroke and then expands. Should 
the piston strike the cylinder head, the shock is absorbed by 
the spring a. In the front head is a stuffing box, the gland of 
which is held in place by the cap h. 

Contrary to the usual design of rotation devices, the rifie-bar 
k is soHdly screwed into the piston, engaging with the rifled nut 
m, in the ratchet g. The outer end of the rifie-bar extends into 
the closed tube r, which is connected with the cyHnder by the 
small passages x and y. Hence, live air acts on the entire rear- 
end area of the piston, including the area of the rifle-bar. The 
tube f, and connecting passages, serve incidentally for the better 
distribution of oil on both sides of the ratchet and rifle-nut. 



294 



COMPRESSED .\IR PLANT 




=3 



3 



V C 



O' 



o 
1-4 

t4 



COMPRESSED AIR ROCK-DRILLS 295 



Operation of Reciprocating Drills 

Air Pressure. The evidence from recorded tests shows 
conclusively that low air pressure is uneconomical. The force 
of the blow and number of strokes per minute fall off, resulting 
in a marked decrease in footage of hole drilled. Though drilling 
in soft rock does not require so high an air pressure as for hard, 
the best results are obtained by pressures of 70-80 lbs. Practice 
has tended toward the use of higher pressures, up to 90 or 100 
lbs. ; but, though more work in some kinds of rock may be done 
by pressures above 80 lbs., the life of the drill is shortened and 
cost of repairs increased. The customary nearly uncushioned 
blow, under heavy air pressure on hard rock, is very destructive 
to the machine, and the bits are dulled sooner and are more 
apt to chip. 

Several important series of tests on air consumption at dif- 
ferent pressures were made on the Rand between 1904 and 1909. 
The last of these tests continued for nearly a year.* The rock 
was red granite, a large block of which was embedded in con- 
crete. A quarry bar was used for mounting the drills. All 
holes were drilled vertically, with abundance of water. Two 
receivers holding 757 cu.ft. were employed, the pressure for 
each run being raised by the compressor to 80 lbs., after which 
the receivers were shut off. One drill at a time was operated, 
the run continuing until the receiver pressure dropped to 70 lbs. 
The drill was then stopped, and the depth and diameter of hole 
measured. Similar runs were made with pressures from 70 to 
60, 60 to 50 lbs., etc. The receiver capacity, in terms of cu.ft. 
of free air, was calculated for each run and pressure, correction 
applied for temperature, and the air consumed based on the 
volume of free air at 70° F. and 24.8 ins. of barometer (cor- 
responding to the Rand altitude of 5,000 ft.). 

Eliminating results of runs indicating erratic behavior of 

* Mech. Engs. Assoc, of the Witwatersrand, 1904 (Abs. in Mines and Minerals, 
Sept., 1904, p. 64); Jour. Transvaal Inst. Mech. Engs., Nov , 1907 and Feb., 1908; 
Eng. b" Min. Jour., Jan 21, and Feb. 18, 1911. 



296 



COIl^IPRESSED AIR PLANT 



some of the drills, due to being in poor condition, a test of 13 
3i-in. drills, with 3-in. bits, gave the following averages: 

Table XXX 







Gage 


Pressure, Lbs. 






80-70 


70-60 


60-50 


50-40 


40-35 


Linear inches drilled per min 

Cu.ft. free air per minute 

Cu.ft. free air per linear in. of hole. . . . 
Ditto per cu in. of hole 


1-3 
124. 

95-3 
133 


I . I 
117. 
106.4 
14.8 


1 .0 

TOO. 

100. 
13-8 


0.6 
70. 
116. 4 
150 


05 
60. 
120. 
16 6 







Each run occupied about 6 mins. Some of the average 
results were not consistent, and individual figures, of course, 
showed still greater variations. These were due to lack of 
uniformity of the rock, differences in temper and sharpness of 
bits and the personal equation of the drill-runners, each of whom 
" was selected by the agent of the maker of the drill." The 
lengthy paper from which these data are taken includes many 
tables, thoroughly summarizing the work. Among other points, 
the importance of the question of air pressure is clearly demon- 
strated. 

Consumption of Air. Due to the irregularity of the work of 
machine drilling, and the fact that a number of drills are generally 
operated by the same compressor plant, few figures are available 
as to the actual air consumption of a single machine. Average 
figures, however, are the only really useful ones. The duty is 
usually based on the consumption of free air per min., which 
depends on the size of drill, air pressure, character of rock, 
and the proportion of the total time actually occupied in drilUng. 
The compressor capacity for one drill is evidently greater than 
the average required for a number of machines. With a large 
number, the delays to which each is subject, for setting up or 
shifting, changing bits, stoppages caused by the bit sticking in 
the hole, etc., make it improbable that all will be in simultaneous 
operation; hence, the average allowance of air for each may be 
reduced. Momentary or occasional peaks in the load on the 



COMPRESSED AIR ROCK-DRILLS 



297 



compressor, when an unusual number of drills are working 
simultaneously, may be disregarded; or at least need not be 
provided for by increasing the compressor capacity. 

Rock-drills of different makers, even when of the same 
diameter of cylinder, vary in their consumption of air and 
reliable figures are not easily obtained. Table XXXI, showing 
the free air per minute required for one drill, is based on a com- 
parison of the statements of several manufacturers, checked by 
recorded tests. It represents, within reasonable limits of error, 
actual practice for machines in good order. No allowance is 
made for the preventable loss of air in leaky pipes, nor for 
frictional loss of pressure in transmission (see Chap. XVI). 

Table XXXI 

Cubic Feet of Free Air per Minute Consumed by One 

Drill at Sea-Level 



Gage 
Pres- 






Diameters of 


Drill 


Cylinder in 


Ins. 
































sure. 


2 


2\ 


2\ 


2f 


3 


38 


016 


3i 


3-1 


3f 


4r 


5 


60 


58 


63 


70 


82 


90 


97 


100 


105 


114 


118 


135 


155 


70 


62 


72 


80 


92 


104 


112 


IIS 


118 


130 


135 


152 


174 


80 


70 


80 


88 


103 


115 


125 


130 


135 


142 


153 


173 


205 


90 


78 


87 


95 


115 


128 


137 


141 


148 


165 


173 


194 


222 


100 


8S 


96 


108 


126 


140 


151 


155 


161 


176 


184 


210 


250 



When a number of drills are operated by the same plant, the 
compressor capacity for furnishing the total average quantity of 
free air required per minute, at sea-level, may be found approxi- 
mately by the following table of multipliers : 

Table XXXII 



Number of drills . . 


I 


2 


1 



4 


5 


6 


7 


8 


9 


10 


Multiplier 


I 


1.8 


2.7 


3-4 


41 


4.8 


5-45 


6.1 


6.7 


7-3 


Number of drills. . . 


. II 


12 


15 


20 


25 


30 


35 


40 


50 


60 


Multiplier 


7.8 


8.4 


10.3 


12.8 


i^i 


173 


19.7 


22.0 


26.5 


30. 5 



298 COMPRESSED AIR PLANT 

To use the tables multiply the cubic feet of free air per 
minute consumed by one drill (Table XXXI) , by the multipher 
corresponding to the number of drills (Table XXXII). 

To Find the Compressor Horse-Power required for any 
number of drills at any altitude. Example: lo 2i-in. drills, 
working at 12,000 ft. altitude, mth air at 80 lbs. gage. 

From Table XXXI, i drill at sea-level uses 80 cu.ft. free air, 

and 10 drills (Table XXXII), 80X7.3 = 584 cu.ft. At 12,000 

ft., as the relative output for 80 lbs. (Table XIII, p. 194) is 

0.65, the compressor capacity is 584-^o.65 = 900 cu.ft. The 

volume of compressed air per I.H.P. at 80 lbs. and 12,000 ft. 

0.02 j; +0.646 , ^ - . , VP 

= =0.785 cu.tt. In the equation V =-^, P, from 

Table XIII, =9.34; hence, V' = ^^^^^^ = 94, and ^"^ 



80+9.34 ^^' 0.785 

119.7H.P. 

The horse-power may also be found by using the ratio of 

compression directly, thus: cu.ft. at sea-level (see above) 

-r^ • r • - 80+9.34 
= 584. Ratio of :ompression at 12,000 tt. = =9-56, 

which (Table VI, p. 141) corresponds to 127 lbs. gage, and the 
H.P. required to compress i cu.ft. to this pressure = 0.204. 
Hence, 584X0.204 = 119.13 H.P., which agrees closely with the 
result obtained by the first method. 

The figures in Tables XXXI and XXXII are not exactly 
applicable to all cases. iModifying factors are: 

(i) Kind of Work. The time required to set up the drill 
depends greatly on the shape of the working, whether a tunnel or 
drift, a shaft, stope, or open cut. If the floor and roof, or the side 
walls, of a mine opening are irregular or loose, much time may be 
lost in shifting the machine and setting it up, according as it is 
mounted on column or tripod. 

(2) Character of Rock also influences the air consumption. 
In hard rock the advance in drilling is slower than in soft, so 
that the machine makes longer continuous runs. Less total 
time is occupied in shifting and setting up for drilling the suc- 
cessive holes of a round, and the consumption of air per unit 



COMPRESSED AIR ROCK-DRILLS 299 

of time is therefore greater. Though this increase is partly 
offset by the fact that the bits are more quickly dulled in hard 
rock and must be changed at shorter intervals; still, in very hard 
ground the drills may be kept running with but few and short 
intermissions. In soft rock, though the actual speed of .drilling 
is greater, there are apt to be more delays due to rifling of the 
hole and sticking or " fitchering " of the bit. On the whole, 
for hard rock it is advisable to provide a greater compressor 
capacity than is given in the tables. The compressor can then 
run at a slower speed, thus avoiding excessive heating of the air. 
The time actually occupied in drilling will vary for each machine 
from, say, 4 to 6 hours out of an 8-hour shift. 

(3) Physical Condition of the Drill is important. The 
tabulated figures are for new machines, or those in good order. 
More air is consumed by old drills, with valves and pistons so 
worn that they do not fit closely. Even for drills in fair average 
condition, this is clearly shown by the fact that the exhaust, 
instead of being short and sharp, is nearly continuous. A large 
allowance must be made for old machines. 

If definite values could be assigned to the above items, 
estimates of air consumption could be made for any given 
condition. Though this is manifestly impossible, a few averages 
for an entire shift's work have been recorded by Messrs. J. E. 
Bell and L. L. Summers. For a 3-in. drill, the volume of free 
air required per shift of 8-hours is as follows in Table XXXIII, 
the gage pressure being 100 lbs. : 

Table XXXIII 



Elevation. 



Sea-level , 

5,000 ft . 
10,000 ft , 



Cubic Feet of Free Air. 



Per Shift of 8 Hours. 



25,000-42,000 
30,000-49,000 
35,000-60,000 



Per Minute. 



52.1- 87.5 
62.5-102.0 
73.0-125.0 



These figures include all deductions for delays and stoppages. 
Taking the various allowances into account, and applying 



300 COMPRESSED AIR PLANT 

them to Tables XXXI and XXXII, the following results of an 
elaborate test made at the Rose Deep Mine, Johannesburg, 
South Africa,* will be found in fairly close agreement with 
what precedes. The average number* of drills (IngersoU-Ser- 
geant), of several sizes, operated during the 6-hour test, was 
calculated to be equivalent to 30.9 3i-in. drills. Average duty 
per drill, 4 ft. 5f| ins. of hole per hour (diameter of hole not 
stated). Average air pressure, 69.83 lbs. Free air used per 
drill per minute, 81.08 cu.ft. It is fair to assume that most of 
the drills were more or less worn, or at least not in perfect con- 
dition. According to the tables, the average free-air consump- 
tion for 30.9 drills should have been 68 cu.ft. per min., or about 
15% less than that shown by the test. This difference is due 
in part to the altitude above sea-level. The compressor horse- 
power per drill was 12.72; but as the work done during the 
6-hour test was approximately equal to that usually accompUshed 
in 8 hours of regular work, the actual horse-power per drill under 
normal conditions in this mine may be taken as 12.72X1 = 9.54. 
The air piping was known to be remarkably free from leaks. 

Another test run, on 75 3 J-in. drills was made at the Champion 
Iron Mine, Michigan.! At 78 lbs. normal gage pressure the 
average consumption for the day shift for i month was 67.1 
cu.ft. of free air per min. The pressure usually dropped consider- 
ably, however, when work was in active progress. According 
to the tables, 75 drills should use about 58.5 cu.ft. of free air 
per min., or 13% less than shown by the test. 

A large part of the compressed air used in the mines of the 
central Rand is now bought from the Rand Mines Power Supply 
Co., on meter measurement at a pressure of about 100 lbs. The 
following notes are from a paper by E. G. Izod and E. J. Lasch- 
inger {Trans. So. Af. Inst. M. E., Oct., 1913). The unit of 
measurement is 440 cu.ft. at 12.1 lbs. (Rand atmospheric pres- 
sure), weighed at 60° F.; it represents 641 watt-hours, which 
is the energy developed by the isothermal expansion of 440 
cu.ft. of air, from 100 lbs. to o lb. gage. 

* L. I. Seymour, South African Association of Engineers, 1898. 
t Eng. and Min. Jour., May 18, 1905, p. 037. 



COMPRESSED AIR ROCK-DRILLS 



301 



A 3i-in. rock-drill in stoping uses an average of 80 air- 
power units per 8-hour shift, equivalent to 35,200 cu.ft. free 
Rand air; on development work, about 70,000 cu.ft. A useful 
curve diagram for checking air consumption is obtained by 
platting the average air units per drill shift as ordinates, and 
the ratio of development shifts to total shifts as abscissas. By 
attention to this, one large group of mines reduced its air con- 
sumption per drill shift from 163 to 138 units. Paying for air 
by meter leads to more care in its use. 

At the Village Main Reef Mine extensive tests were made in 
1913, as summarized in Table XXXIV. They show the impor- 
tance of a close fit of the piston in the cylinder. 

Table XXXIV 
Tests on a 3.183-lNCH Ingersoll Drill 



Gage Pressure. 



Piston diam., ins 

Difference between cyl. bore and piston diam., 

ins 

Strokes per min 

Lbs. free air per min 

Lbs. air per double stroke 

Cu.ft. free air per min. at 68° F. and 12.2 lbs. 

absolute pressure 



70 Lbs. 


90 L 


3-177 


3-157 


3-177 


0.006 


0.026 


0.006 


236 


235 


250 


6.9 


9-4 


10. 2 


0.0292 


0.04 


. 0408 


112 


152 


165 



3-157 

0.026 
262 

12.8 

0.049 
207 



Efficiency. Though it is well-known that compressed-air 
drills are uneconomical in consumption of power, it is difficult 
to reach definite conclusions as to their efficiency. The useful 
work in the ordinary mechanical sense, done by a drill in making 
a hole of given depth and diameter in a rock of given hardness, 
toughness, and general physical character cannot be determined 
absolutely. All that is known is that the drill requires a cer- 
tain volume of air per minute, which has been furnished by the 
expenditure of a certain average horse-power at the compressor. 
Comparative figures of work done, in terms of speed of drilling 
per cubic foot of free air consumed, are useful as far as they go, 
and are the only practical basis for estimating efficiency. But 



302 COMPRESSED AIR PLANT 

the results obtained do not accurately represent the efficiency 
of machine drills as compared with other air or steam 
engines. 

In their operation, rock-drills differ greatly from other com- 
pressed-air machines, because the personal element of the skill 
and experience of the drill-runner exerts so important an influence 
upon the amount of work accomphshed, and because the rate of 
drilHng is so greatly modified by the physical and mineralogical 
character of the rock, together with the purely adventitious 
occurrence of cracks, slips, and fissures. A skillful drill-runner 
will inevitably do more work per shift than an inexperienced 
man, and will make a faster advance in rock with which he is 
familiar than in rock that is new to him. 

Therefore, though mechanical efficiency is the basis upon 
which machines in general are compared, in the case of com- 
pressed-air drills it is not the only consideration, nor is it the 
most important. Their efficiency is subordinate to the attributes 
of strength, simplicity, portabihty, durabihty, facility with 
which repairs may be made and capacity for work in terms 
of depth of hole drilled per unit of time. They must withstand 
hard and often unintelligent usage. The strong point of com- 
pressed-air drills is their ready applicability in their special 
field of work. In possessing a cylinder, piston, and valve, the 
drill roughly resembles a steam engine, but there the likeness 
ceases. Severe shock and vibration are essential accompani- 
ments of its work. No fly-wheel is admissible, or other means 
of storing up and equalizing the power. The service demanded 
is therefore totally different from that performed by ordinary 
engines. 

The low mechanical efficiency of the rock-drill is due mainly 
to the fact that air is admitted to the cylinder practically 
throughout full stroke. Hence, the valve motion resembles 
that of many simple, direct-acting pumps. Expansive use of 
the air to any extent is not practicable, both because of the 
undesirabiHty of introducing complexity of mechanism in 
machines subjected to rough usage and the difficulty of adapting 
cutoff gear to the variable length of stroke required. The drill 



COMPRESSED AIR ROCK-DRILLS 303 

cannot be kept always at full stroke; while in operation it is 
often necessary to feed the machine so far forward that the 
length of stroke is no more than i in., and the valve must still be 
able to cause a sharp, quick reversal of the stroke. The useful 
work is done on the forward stroke, in striking the blow. If the 
valve be thrown too soon, the stroke of the piston will be 'short- 
ened; if too late, the piston will strike the cyHnder head. Rock- 
drills, therefore, cannot attain the economy resulting in other 
air motors from using the air expansively. Incidentally, using 
air at full stroke is of some advantage, because exhausting at 
high pressure in a measure prevents troublesome accumulation 
of ice, in case the air is moist. Freezing, if any, is at least con- 
fined to the outer portion of the exhaust port, whence it is easily 
removed. 

In dry, dusty mines, tappet- valve drills give good service. 
When a drill is not in use, and disconnected from the air hose, 
dust and grit may enter through the ports, passing thence 
into the valve chest and cylinder on resuming drilling. The 
wear and consequent looseness in the fit of the moving parts 
thus caused is apt to have a more unfavorable effect on the 
operation of the spool than the tappet valve. Leakage of air 
past the valve or piston prevents proper action of the auxiliary 
ports, not only producing irregularity in reversal and shortening 
of the stroke, but diminishing the drill's efficiency. It is true 
that the tappet valve involves the use of one extra part and, 
in case of the pivoted three-arm tappet, breakage is not infre- 
quent. But while the spool valve is strong and reliable, its 
maintenance cost in dusty mines is higher than that of the 
tappet-valve motion. 

The maximum force of blow is attained by drills working 
without cutoff on the forward stroke, and the best drills are 
thus designed. The valve is not reversed until the blow is 
delivered, and the exhaust is free, with but Httle back-pressure 
on the piston. Cushioning was formerly a feature of some drills, 
with the idea of reducing shock, but it is now recognized that 
an uncushioned blow gives greater efficiency. A drill so designed 
may strike too heavy a blow in very hard rock, the remedy 



304 COMPRESSED .\IR PL.\NT 

being to feed the drill-head down, so as to work with shorter 
stroke. 

On the back stroke cushioning is desirable, to ease the rever- 
sal and prevent injury due to the piston striking the cyhnder 
head. The back-stroke cushion is produced by cutting off the 
exhaust before the end of the stroke. Only enough power needs 
to be developed on this stroke to overcome the resistance due to 
the weight of the moving parts, and the frequent tendency for the 
bit to stick in the hole. 

In ordinary machine drills, the piston speed should not be too 
great — say, not much over 350-375 strokes per minute. Relative 
speeds of stroke do not constitute a proper basis for the com- 
parison of efhciencies. To give effect to the blow, the weight 
of the moving parts must be relatively great, and very high 
speed is attended by excessive wear and breakage. These con- 
clusions do not apply to drills of the hammer type (Chap. 
XXI), which strike a light blow at a high speed; the weight 
of the mo\dng parts is small. 

Drill Repairs. In choosing a driU the question of repairs 
is of great importance. But little useful information concerning 
this point is available; such data could be obtained only by 
operating drills of several different kinds under the same con- 
ditions, and for a considerable period of time. Such oppor- 
tunities are rare. It is generally inadvisable to use different 
makes or more than two sizes of drill in the same mine, since 
this requires the keeping of duplicate spare parts for each. 

The item of repairs depends largely upon the experience and 
character of the drill runner. A careful man treats his machine 
wdth intelligent consideration. He will set it up properly, to 
reduce the risk of getting out of ahgnment as the hole is deepened; 
and, if the bit should stick {" fitcher ''), he keeps his temper and 
refrains from striking unnecessarily heavy blows on the drill head 
or chuck. A fitchered bit may often be loosened by slacking the 
clamp bolts, thus allowing the machine slightly to shift its posi- 
tion. The serious abuse to which machine drills are frequently 
subjected may be reduced by efficient supervision on the part of 
foremen and shift-bosses. 



COMPRESSED AIR ROCK-DRILLS 305 

In every machine there should be as few moving parts as 
possible. But the work of a rock-drill is severe, and it is often 
operated by incompetent men. Even when run with care, 
wear is rapid and breakages are frequent. The maker therefore 
has the problem of producing a drill comprising as few parts as 
practicable, and designing it so that the parts especially hable 
to wear or breakage may be replaced readily, cheaply and 
without the necessity of discarding a larger piece, with which 
the broken part may be connected. New cyHnders may be 
bored out to fit worn pistons, or new pistons fitted to old cylinders. 

When new, the piston should have the closest possible work- 
ing fit. The difference between the cylinder bore and piston 
diameter should not exceed ^-in. Modern drill cylinders are 
counterbored at each end, to facilitate reboring. The difference 
in diameter between a spool- valve and the bore of its chest should 
not exceed rJo-in. Chuck and chuck keys must be in good 
order, and the bit when set must be in accurate alignment with 
the piston. A well-equipped repair shop lengthens the Hfe 
of machine drills. Drill runners should not be encouraged to 
tinker their machines underground. If repairs or adjustment 
be necessary, the drill should be sent at once to the shop.* 

Repair Costs. While no • generalizations are possible, on 
account of the extremely variable service of rock-drills, Tables 
XXXV and XXXVI, C. K. Hitchcock, Jr. {S. of M. Quarterly, 
1914), of repairs costs at a Michigan copper mine, will be found 
significant and useful. 

The records of cost of labor and material (not including 
mountings nor connections) were kept by the month, the cost at 
the end of each month being platted on a chart against the 
number of days, of two shifts, that the drill had worked since 
being put into commission. From the chart the costs for each 
period of 25 days were scaled, as recorded in the tables. Most 
of the smaller repair parts were made in the mine shops, the 
larger parts being bought from the makers. Table XXXV 

* For a detailed discussion of maintenance and repairs see Chap. VIII, " Rock 
Drills," by E. M. Weston. See also Min, and Sci. Press., 1905, Nov. 4,' p. 308; 
Nov. II, p. 329, 



306 



COMPRESSED AIR PL.\XT 



shows that the average costs between loo and 150 days of senice 
were about $7.00 for each 25 days. 

Table XXXV 
Repair Costs of 15 New 2f-lNCH Drills 









Days 


m Use, 2 Shifts per Day. 




Drill. 


















25 


50 


75 


100 


125 


150 


175 


I 


$ 0.60 


S 1.25 


S 1.50 


$ 2.40 


$ 300 


$ 7 05 


S10.80 


2 


3-75 


7 50 


7.90 


8 


75 


12.00 


16.25 


21:. 80 


3 


3 30 


10.00 


II. 10 


13 


50 


20.30 


40.75 


40.91 


4 


.90 


I 25 


4 55 


5 


30 


28.39 


28.39 


28.39 


5 


•15 


.40 


I 50 


4 


50 


580 


6.40 


9. 20 


6 


3.80 


7.40 


8.00 


8 


08 


8.08 


12.15 




7 


.20 


•50 


1. 10 


4 


75 


8.00 


14 -95 


16. 20 


8 


■55 


I 05 


5.60 


5 


91 


7.80 


10.57 


10. 70 


9 


4.00 


6. 20 


8.70 


12 


30 


41. 20 


54-59 




10 


.70 


.70 


350 


4 


70 


8.29 


8.45 




II 


.60 


1 .00 


I 50 


21 


00 


41.00 


42. 10 


43 36 


12 


•55 


1 .00 


2. 20 


9 


10 


12.00 


12.30 


21.30 


13 


•85 


1 . 20 


4.40 


7 


70 


8.70 


10.25 


18.14 


14 


3-65 


4. 10 


4-85 


5 


05 


8.50 


50.50 


52.01 


15 


2.55 


4-35 


465 


5 


20 


II . 20 


14.60 


14-75 


Totals. . . . 


$26.15 


S47.90 


S78.05 


$118.24 


S224. 26 


5329.30 




Averages . . 


1-74 


1 3.:9 


4 74 




58 


14 95 


21.95 





Table XXXVI indicates a repair cost of about S5.00 per 25 
days. In the upkeep of machine drills, both economy and 
efl&ciency can be improved by discarding drills having abnormally 
poor records, or by overhauling them so thoroughly as to cut 
down cost of maintenance. 

At a large group of mines on the Rand, the complete repair 
cost of standard piston drills (including air hose) averaged S42.00 
per 52 shifts. On introducing a contract system of drill repairs, 
the cost was reduced to $27.00, and it is thought it might be 
further reduced to S21.00 or S23.00. 

Records of Work. Referring to the preceding discussion of 
the operation of machine drills, if approximately complete 
records of work could be secured, useful tabulations could be 
jii^e of the comparative speeds of drilling in different rocks 



COMPRESSED AIR ROCK-DRILLS 



m 



Table XXXVI 
Repair Costs of Six 3|-Inch Drills 





Number of Days in Use. 


Drill. 


25 


50 


75 


100 


125 


i5o 


A 
B 
C 
D 
E 
F 


$ 2.00 

3.00 

6. 10 

4. 20 

28.00 

-50 


$ 2.50 
11.30 
14. 20 

5.20 
33 30 

4-30 


$ 4.00 
19. 10 
18.20 

6. 20 
36.10 

4.90 


$ 5-40 
22.30 
19.60 
9. 20 
37.00 
11. 00 


$ 8.00 
27.00 
24.50 
18.10 
38.10 
12.80 


$ 19.00 
34-00 
28.60 
18.20 

41-50 
14.00 


Totals 

Averages 


$ 43-80 
7-30 


$ 70.80 
11.80 


$ 88.50 
14-75 


$104.50 
17.42 


$128.50 
21.42 


$155-30 
25.88 





Number of Days in Use. 


Drill. 


175 


200 


225 


250 


275 


300 


A 
B 
C 
D 
E 
F 


$ 22.00 
40. 20 
28.90 
18.30 
44.70 
17.80 


$ 22.60 
41-30 
28.90 
21.50 
45-90 
18.20 


$ 40.00 
42. 20 
32.60 
22. 10 
46.50 
18.20 


$ 64 . 20 
46.70 
33-40 
46.00 
46.90 
35 00 


$84.80 
53 20 
46.30 
46.00 
48. 20 
55-20 


$ 88.10 
59.60 
49.20 
47-60 
49-50 
59 30 


Totals 

Averages 


$171.90 
28.65 


$178.40 
29 -73 


$201.60 
33 -60 


$272.20 
45-37 


%2>2,3> - 70 
55-62 


$353-30 
58.88 





Number of Days in Use. 


Drill. 


325 


350 


375 


400 


425 


450 


A 
B 
C 
D 
E 
F 


$ 97 00 
66.00 
50.20 
48. 20 
72.00 
60 . 20 


$ 99-50 
66.80 
48.80 
53-70 
90-50 
60. 70 


$101 . 50 
67.00 
54- 50 
54.60 
III . 20 
63.00 


$102.00 
78.80 
61.00 
57-10 
1 1 2 . 00 
63-70 


$115. 20 
79 50 
75-00 
60. 20 
113.80 
64.00 


$76.70 
62. 70 

113.80 
70.00 


Totals 

Averages 


$393 • 60 
65-60 


$420.00 
70.00 


$451-80 
75-30 


$474.60 
79.10 


$508.00 
84-67 





308 COMPRESSED AIR PLANT 

and ores. Having such data, it might even be possible to 
designate the kinds of service for which the different makes 
and t}^es of drill are best adapted. But, rock and ore charac- 
teristics, and other local conditions, vary so greatly that detailed 
comparisons are of doubtful value. The results of " test runs " 
are sometimes cited to show that one machine is better than 
another, in the sense that it can drill faster or that it uses less 
air for the same footage. By operating drills of dift'erent type 
side by side, in the same rock and with the same air pressure, 
some approach to an accurate comparison would be possible; 
but, even then, it is clear that the physical condition of the 
competing drills, and the " personal equations " of the respective 
drill runners, aft'ect the rate of work and cost per foot of hole. 
Moreover, as test runs are made under the stimulus of rivalry 
(and perhaps with a " bonus " to the winner), they do not con- 
stitute a useful basis for ascertaining the relative merits of 
dift"erent drills, nor the footage that could reasonably be expected 
in ordinary daily work. For further elucidation of these con- 
siderations the reader is referred to the voluminous records of 
the elaborate drill tests made on the Rand (see footnote, p. 295). 
In recording the work of machine drills, the important items 
are: 
(i) Kind of rock or ore. 

(2) Type and size of drill. 

(3) Air pressure at the drill. 

(4) Diameter of hole. 

(5) Total elapsed time for a hole of given depth. 

(6) Measured volumes of free and compressed air consumed 

for the depth drilled. 

(7) Time occupied in setting up, changing bits, and delays due to 

breakage or derangement of the drill, or to '' fitchering " 
of bits. 

(8) Net drilling time per hole and per foot or inch of hole, 

resulting in the number of inches per minute (or feet per 
hour) drilled while the machine is in actual operation. 
Examples of Drilling Speeds. Table XXXVII gives figures 
based on a number of recorded runs. 



COMPRESSED AIR ROCK-DRILLS 



309 



Table XXXVII 



Kind of Work. 



Crosscut, 5 X 7 f t . . 
Crosscut, 5X7 f t . . 



Drift. 

Tunnel, 7x7 ft. 
Stoping 



Stoping , 
Drift..., 



Stoping and drifting. 

Stoping 

Stoping 

Stoping 



Stoping 

Crosscut 

Stoping 

Stoping 

Stoping 

Drift, 10X10 ft 

Stoping 

Drift, 4-5X6 ft. 
Stoping 



Drift. 



Drift. 



Stoping 

Drift 

Crosscut, 9X9 ft 



Rock or Ore. 



Hard quartzite and 

limestone. 
Hard quartzite and 

limestone. 

Quartzite 

Basalt 

Amygdaloid copper 

rock. 
Amygdaloid copper 

rock. 
Amygdaloid copper 

rock. 

Hard limestone 



Hard limestone 

Magnetic iron ore. . . . 
Rather soft porphyry, 

etc. 
Very hard quartzite . . 
Quartz and hard slate. 

Hard quartz 

Hard hematite 

Quartzose 

Hard limestone 

Pyritic ore 

Limestone 

Amygdaloid copper 

rock. 
Amygdaloid copper 

rock. 
Amygdaloid copper 

rock. 
Hard phonolite breccia 
Hard phonolite breccia 
Quartzite 



Size of 
Drill, 
Ins. 



3t 

3 

3i 



3l 



3t 



\2f / 



9^ 
^4 



2f 



3 

2| 

3i 
3i 

2f 

3&^3l 
3&^3i 



Air 
Pres- 
sure, 
Lbs. 



60 



60 



63 



63 



65 



75 
80 

75 



75 
73 
73 
74 



105 



Aver. 

Depth 

of Hole, 

Ft. 



3 

3f 

5h 



61 
6^ 
4l 

3i 



7 

5i 
4 
4 

H 



si 
5! 



4t 

3 3 



Inches of Hole 
per Minute. 



Total 
Time. 



0.43 
0.40 

1 . 29 

0. 72 
0-93 

1 .60 

1 . 26 

1 .00 



I-3I 
1.38 

0-54 
1.50 
1 . 20 
1.28 
I-5I 
1-35 

1 . 20 

2. 62 
1.67 

1.46 

1. 61 

0.89 
2. 60 
1. 17 



Net 
Time. 



1-55 

2.30 
2. 70 

2.46 



2.25 



2.45 
2-34 
1.97 

0-95 
3.38 

2.43 



Conclusions. The average speed of drilling based on column 
6 of Table XXXVII is 6.4 ft. per hour. In general the duty 
of a standard 3-in. drill, in rock or ore of average hardness, 



310 COMPRESSED AIR PLANT 

ranges from 40-50 ft. per 8-hour shift, including all ordinary 
delays for setting up and changing bits. For very hard, tough 
ground, the speed is often lower, while much more than 50 ft. 
per shift may be made when the conditions are favorable, and 
also in drilling deep holes, for which fewer set-ups are required. 
The cost per foot of hole is extremely variable, ranging from say 
8 cents in easy ground, and where wages are low, up to 25 cents 
under adverse conditions. 

For moderately soft ground, not requiring holes of large 
diameter to contain the necessary quantity of powder, the smaller 
sizes of machine drill (2-2 J in.) are usually preferable. Their 
first cost and air consumption are less than for large drills, and 
they may be operated by one man. Small machines are espe- 
cially useful for stoping in thin veins. For hard ground, and 
as a rule in shaft sinking, tunnelling, cross-cutting and similar 
work, the 2f, 3, and 3i-in. sizes are best. For deep holes in 
open-cut work, still heavier drills are often necessary — up to 
3 J in., or even larger. 

Hammer drills (Chap. XXI) are now successful competitors 
of reciprocating drills for nearly all kinds of rock excavation, 
including many of the operations of mining. 



CHAPTER XXI 

HAMMER DRILLS 

The principle of the hammer drill was first applied in pneu- 
matic riveters, and tools for chipping, rough chiseling and mis- 
cellaneous machine shop work. Their earliest employment in 
mines was for cutting hitches for timbers, blockholing, and other 
shallow drilling. In recent years they have rapidly grown in 
favor. They have largely displaced reciprocating machines 
for s toping and similar work, and are often used for sinking 
shafts and winzes. In other words, they are best applied to drill- 
ing holes directed steeply upward or downward. For tunnel- 
ling, drifting, cross-cutting, etc., reciprocating drills are still in 
general use, though hammer drills are employed to some extent 
in these operations also. 

Classification. The author proposes the following classi- 
fication, as being more useful than one based upon the type 
of valve-motion: 

A. Machines designed for tripod or column mounting; 
usually of the larger sizes, comparable to standard drills of the 
reciprocating type. 

B. Small machines with either a cross or D-shaped handle; 
chiefly for drilHng holes directed steeply downward, the weight 
of the drill resting on the bit itself. 

C. Machines with an automatic air-feed standard; primarily 
for overhead stoping, though they may also be mounted on a 
column, as for breast work. 

General Construction. In the hammer drill, the bit does not 
reciprocate; its shank projects into the forward end of the 
cylinder and is struck a rapid succession of blows by the piston, 
which acts as a hammer. The cutting edge of the bit is in 
constant contact with the bottom of the hole, except during the 

311 



312 COMPRESSED AIR PLANT 

slight rebound caused by each blow of the piston. Some of 
these machines, like the Hardsocg, are valveless, the functions 
of the valve being performed by the reciprocations of the piston. 
Others, hke the Leyner-Ingersoll, Sullivan, and Climax, have 
air- thrown valves. At first, no attempt was made to introduce 
automatic rotation of the bit; the operator simply turned the 
drill back and forth on its axis, by means of the handle. At 
the present time, most hammer drills are provided with rifie-bar 
rotations, similar to that of the reciprocating machines, or some 
modification of that device. The bit shank is octagonal, gen- 
erally fitting loosely in a corresponding chuck socket. To keep 
the hole round and reduce the chances of rifling, the bit is com- 
monly of the star or rosette shape, with 6 (sometimes 8) radial 
cutting edges. The cutting edges are thus so close together that 
even if several successive blows are struck in the same angular 
position of the bit, the ridges of rock between the edges are 
broken away. 

As the bit does not reciprocate, it is evident that, for down 
holes more than a few inches deep, some automatic means must 
be provided for removing the drill dust or sludge and keeping 
the bottom of the hole clean ; otherwise much of the useful effect 
of the hammer blows would be lost. To accomplish this, a 
hollow bit is generally used, a small hole being bored longitudi- 
nally through its center. By injecting a jet of water, the drilhngs 
are displaced and the bit is kept cool. The same result is 
attained by a jet of compressed air, which produces a low tem- 
perature on expanding. As the speed of stroke is great, the 
cooling of the bit in dry holes is important. The water jet is 
usually applied as described under the Sullivan " Liteweight " 
drill, Chap. XX, p. 278 (see also the " Leyner-Ingersoll drill, 
below). When an air jet is used, it is delivered through a 
similar axial tube in the drill, the rear end of the tube being 
connected with the air-feed pipe; or a small quantity of com- 
pressed air may be led directly to the hollow bit from the valve 
chest. 

The small hand hammer drills (class B), used chiefly for 
down holes, are fed simply by keeping the bit pressed firmly 



HAMMER DRILLS 313 

against the bottom of the hole. Usually the automatic feed, 
for stoping drills (class C), consists of a light telescopic standard, 
attached to the drill head. It is supplied with compressed air, 
which keeps the drill fed up to its work as the hole is deepened. 
Incidentally this device furnishes an air cushion, relieving the 
operator handling the drill from much of the annoyance caused 
by vibration. For breast-stoping, drifting, and other horizontal 
work, these machines may be mounted on a light column. 
Examples of each class of hammer drills are given below. 

Class A. Large Hammer Drills 

Leyner-IngersoU Water Drills, Nos. i8 and 26 (Figs. 150, 
151), are provided with the "Butterfly" valve, for an illus- 
trated description of which see Ingersoll-Rand "Butterfly" 
reciprocating drill (Chap. XX, p. 283). 

Rotation of the bit is effected as follows (see also Fig. 152): 
The rifle-bar i, which is provided with a ratchet and three 
pawls, engages with a rifle-nut, screwed into the hollow rear end 
of the hammer 2. This causes rotation of the hammer on each 
back stroke. The forward, smaller end 3 of the hammer is 
fluted, and engages with an internally fluted bronze nut in the 
rear end of the chuck. Thus the chuck, holding the bit, is 
caused to rotate with the hammer. 

The drill bit is hollow, for the passage of the water into 
the drill hole. The bit for the No. 18 drill has two lugs by 
which it is locked in the chuck. Fig. 153 shows the bit shanks 
of both drills. Another model of the No. 18 drill has an " anvil 
block," inserted between the hammer and bit. The steel has 
no lugs, but fits in a bushing screwed into the chuck, and the 
hammer strikes the anvil block, instead of the bit. In hard 
ground, which does not " ravel " (run into the hole on with- 
drawing the bit), this design works well. By omitting the lugs, 
there is a saving in blacksmithing, and, as the drill-runner 
cannot back the machine out of the hole with full air pressure 
on, breakage of parts is reduced. The smaller drill. No. 26 
(Fig. 152), has a different chuck, designed for bits formed with 



314 



COMPRESSED AIR PLANT 




C50 



o 



«$ 

^ 



£3 



(U 


>> 


— < 


(4 




« 















o 


n 


to 


\ 


M 




6 




l-( 




fe 





CID 



vO 



o 



c3 

o 












/%r 



^5 



HAMMER DRILLS 



315 



a collar and hexagon shank. To retain the bit in the chuck, a 
pin 7 is dropped into a hole in the front head. The weight 
of the drill only is 95 lbs. 




Fig. 152.— Rotation Device of Leyner-IngersoU Drill, No. 18.* 

The water supply is furnished under pressure from an i8-gal. 
steel tank, weighing 70 lbs., accompanying the drill and con- 
nected to it by a length of hose, as shown in Fig. 136, Chap. XX. 




V/s DRILL SHANK FOR n3 



18 DRILLS 



K— P4^^l 




^\^ 



> 



-3^ 



'^>^" 



I I 






:tm 




Ya DRILL SHANK FOR No 26 DRILLS 
Fig. 153.— Details of Bit Shanks. 



Another hose conveys compressed air from the main to the tank. 
Water is thus forced from the tank through the water tube 5 
(Fig. 150), which passes through the rifle-bar and hammer, 

* The smaller details of design of the present rotation mechanism differ some- 
what from those shown in this cut, but the principle is unchanged. 



316 COMPRESSED AIR PL.\NT 

in the axis of the machine, and is delivered into the hollow bit. 
Air is mixed vdih the water, as the hole is thus cleaned more 
effectually than by water alone. An air jet, without water, 
makes too much dust.* The air comes from the drill when in 
operation, not from the tank. 

Lubrication. In the No. i8 drill, an oil chamber is cast 
under the cyhnder bore. It is filled through a plugged opening 
on top of the cyhnder. There is a small port between the oil 
chamber and cyhnder, so that, as the piston travels over this 
oil port, the oil in the chamber is alternately under Hve air 
and exhaust pressure. A small quantity of oil is thus fed to 
the cyhnder at each stroke. The Xo. 26 drill has an oil cham- 
ber 9 (Fig. 151), on top of the cyhnder, pro\'ided ^vith a 
patented oiler feeding into the valve chest. 

The latest (191 8) Leyner-Ingersoll drills are the Xos. 148 
and 248 (Fig. 153a). Though there are shght modifications in 
construction, the rotation and water features are as in the No. 
18 drill. These machines are identical, except that Xo. 148 
has a light shell of 24-in. feed only, while X^o. 248 may be had 
with either 24 or 30-in. feed. They are valveless as to ad- 
mission and exhaust, the hammer being of the differential 
type; but there is a valve in the cylinder to ehminate back 
pressure and permit an uncushioned blow. Lubrication is by 
the automatic "Heart-beat'' lubricator, connected by a smaU 
port with the rear end of the cyhnder, and operated by the 
air pulsations. Its oil-carr}'ing "cartridge" is recharged as 
necessary. Weight of drihs: Xo. 148, 148 lbs.; X^o. 248, 
156 lbs. 

Sullivan '' DR-6 " Drill (Fig. 154) is mounted on a column 
or tripod, and is designed primarily for drifting and tunnelhng. 
Cylinder diameter, 2| ins.; net weight, 148 lbs.; diameter of 
drill steel, ij in. 

* In late years the dust question has received much attention. Though dust 
from shales, coal and other non-siliceous material is comparativeh^ harmless, that 
coming from siliceous rock or ore is distinctly dangerous, causing " silicosis " or 
miner's consumption. For a discussion of this important subject see Mining 
Engineer's Handbook (Peele, 1918), pp. 1379, 1402. 



HAMMER DRILLS 



317 



00 

C 

00 



D 






A 



1-1 



45 



tuD 









'^- 



C/3 



VO 






O 



o 

l-l 



318 COMPRESSED AIR PLANT 

The hammer a has a short tail piece b, which runs inside of a 
hollow shell valve c, having end seats and placed in the rear of the 
cylinder. The rotation ratchet e is set in the forward end 
of the drill. A long front extension d of the hammer, which 
serves as a rifle-bar, passes through the ratchet sleeve, and 
strikes directly on the drill shank. Straight and spiral grooves, 
/ and g, are cut in the rifled striking end d of the hammer; the 
straight grooves engage with guides in the retaining bushing i, 
which incloses the drill-shank bushing h, and the spiral grooves 
engage with projections in the ratchet sleeve. There is no 
rotation of the bit on the forward stroke. On the back stroke 
the ratchet causes the hammer to ride up on its spiral fluting, 
thus rotating it; and since the hammer, while reciprocating, 
engages \\ith bushing i by means of the straight grooves, the 
chuck bushing h and the bit are also rotated. Like the Leyner- 
Ingersoll bit, pre\dously described, the bit shank has two lugs 
;, fitting in the chuck, to provide the grip for transmitting the 
rotation. 

This machine has a water attachment like that used in the 
*' Liteweight " drill (Chap. XX). It comprises a combined 
water and air jet, a single throttle valve controlHng both air and 
water. Air alone may be used for shallow holes. 

Stephens' ** Climax Imperial " Hammer Drill, made at Carn 
Brea, Cornwall, has a i|-in. cyHnder, weighs 75 lbs., and is 
mounted on a light column or bar. In several features of its 
design this machine differs greatly from xA-merican drills. 

The valve motion (Fig. 155) resembles that of the CHmax 
reciprocating drill (Chap. XX). Air enters by the combined 
air and water tap (detailed section) on the side of the valve 
chest; thence passing by the annular recess b in the piston 
valve a, through c and c\ to the main cyUnder ports //,//'. The 
recesses d,d\ in the valve, communicate with the main exhaust 
(not shown). Air is constantly admitted to both ends of the 
chest by a small groove /, the valve being thrown by exhausting 
through the much larger auxiHary ports e and /.* Ports / are 
alternately brought into communication with the annular recess 

* The device is similar to that in the " Sergeant " drill, Fig. 131, Chap. XX. 




Clamp Swivel and ; 
Cradle Sliding Piece ^ 




-7 — r v-- . ---.-^ , --^-^ ,^ 

f^ Waterway 



SECTION ON LINE A-B 




:^- 



E 



S 



E 



Waterway f com Air 
and Water Tap combined 



Water Regi 

--E 



cz: 



r\rr= 



/!_/ 



^^Av^ 



Airway to Automatic 
FeeiEiston from lap L 



^ 



EL 



SECTION GW LINZ C-D 




rc. T55. — Stephen 




imax Imperial " Hammer Drill. 



To face page ji8. 



HAMMER DRILLS 319 

b, thus releasing the air, by way of ports 5 and s', to the main 
exhaust. The ports / are Hned with hollow, conical plugs g, of 
composition metal, shaped below to the curve of the hammer; 
to prevent leakage of air, they are kept in close contact with 
the hammer by the pressure of the valve-chest, when bolted in 
place. When the plugs wear too loose, a thin washer is inserted 
above them. 

The water for the drill hole is best supplied by gravity, under 
a pressure of say 15 lbs. It enters the combined air and water 
tap, or throttle, through the passage k, to the transverse port / 
in the anvil block u which serves also as the drill holder, or 
chuck. Thence the water passes to the hollow bit (see the 
elevation and the ''section on line AB"). The drill may also 
be used for ''dry" holes, the dust being allayed by an external 
spray from the throttle.* 

The machine has an automatic air-feed cylinder. A small 
piston p, with its rod, is rigidly bolted to the lug q, on the cradle 
r. Air from the throttle passes through passage j to the feed 
cylinder, forcing the drill head forward on the cradle and keep- 
ing the bit pressed against the bottom of the hole. After the 
machine has been fed forward 14 ins. (the working length of 
feed), the air is shut off and the transverse bolt, shown in the 
plan of the cradle, is slacked. The operator then sHdes the 
cradle forward on its support under the machine, tightens up 
the bolt on the serrated edge of the cradle, and proceeds with the 
drilling. Thus, a total feed of twice the length of the cradle — 
or about 28 ins. — is obtained without putting in a longer bit. 

Rotation of the bit is effected by hand. The bit is held 
by friction in the conical socket of the chuck. Gear teeth are 
cut on the periphery of an enlarged part of the chuck, engaging 
with which is a smaller gear n (see general plan and the "section 
on Hne CD"), keyed on a spindle passing to the rear of the 
machine and rotated by the handle m. 

Excellent driUing records have been made by this machine, 
both in England and South Africa. 

*This ''dust qllayer " is described in Ghap. XX, under Climax Drill, 



320 



COMPRESSED AIR PLANT 



Class B. Small Hammer Drills, with Handle 

Hardsocg Wonder Drill* (Figs. 156-7) isa valveless machine, 
made in two types. Fig. 156 shows the older form, with D-han- 
dle for hand rotation. There are 4 sizes, weighing 12, 17, 20 
and 30 lbs., all using hollow steel and employed chiefly for 
down holes. 

Air is admitted at the nipple 2 (to which is attached the 
throttle and hose) and, entering the annular recess 3, acts con- 
stantly on the shoulder 13 of the hammer. On beginning the 
forward stroke, the ports 5 and 6, through the head of the 
hammer, are opposite recess 3, and admit live air into and 
behind the hollow hammer. Since the area thus presented to 

air pressure is much greater than the area of shoulder 13, the 

11 




■^sL:' 



Fig. 156. — Hardsocg Wonder Drill, with D-Handle. 

hammer is driven forward. Just before the hammer strikes the 
bit, ports 5, 6 reach the annular recess 8; through this recess, 
part of the air in and behind the hammer which has caused the 
forward stroke, goes to the exhaust port 4, and part through the 
hollow bit to the bottom of the drill hole. The exhaust having 
taken place, the back stroke is made by the constant pressure 
of the inlet air on the annular shoulder 13 of the hammer. 

This drill can be converted into a stoping or drifting machine 
by remo\ing the large plug and handle 12, and screwing into 
the rear head of the cyhnder an air-feed standard, somewhat 
similar to that shown by Figs. 178, 181 and 183. For making 
breast holes, the air-feed machine is mq^nted on a light column. 
Weights, unmounted, 35 and 87 lbs. 

♦ This drill was one of the earliest of the Hammer drillg, 



HAMMER DRILLS 



321 



Fig. 157 shows a recent form of the Hardsocg drill, with 
automatic rotation: a is the ratchet box, with spring-con- 
trolled pawls; h the rifle-bar, engaging the rifle-nut c in the rear 
end of the hollow hammer d. The hammer is thus rotated like 
the piston of a reciprocating drifl (Chap. XX). The fo;rward 




Fig. 157.— Hardsocg Drill, with Automatic Rotation. 

end e of the hammer has a square cross-section, reciprocating 
in the square bushing/, which holds the octagon shank g of the 
bit. Thus the rotation of the hammer is communicated to 
the bit. A yoke-shaped bit retainer //, somewhat resembling 
that in Fig. 158, holds the bit in its socket. 

Live air enters port i, from the throttle j to the annular 
recess k^ and, as shown by the arrows, passes alongside the 



322 COMPRESSED x\IR PLANT 

rifle-bar, through passages /,/, in the rifle-nut, to the rear of the 
hammer. On completing the forward stroke, the air in the 
rear end of the cyhnder passes to the exhaust port m. When 
the hammer is in this position, the small shoulder n is opposite 
the annular recess k, thus admitting air to act on the large 
shoulder o of the hammer, and causing the back stroke. At the 
end of the forward stroke, some of the exhaust air passes down 
alongside the square head of the hammer (which makes a loose 
fit in bushing /) , and goes through the hollow steel for cleaning 
the drill hole. If required, by turning the throttle 7 backward, 
live air can also be deUvered through the bit, by way of the port 
p. Weight of drill, 50 lbs. 

Murphy Drill is another example of valveless machine, 
made in both the D-handle and air-feed types, by C, T. Carnahan 
Manufacturing Co., Denver, Colo. 

Ingersoll-Rand ' * Jackhamer " is made in two styles: 
'^BCR-430," for dry holes; '^ BCRW-430," for wet holes. 
Referring to Fig. 158, a is the hammer, b the ratchet, c the rifle- 
bar, d the retaining bushing for the hexagonal drill shank. The 
forward end e of the hammer has straight longitudinal fluting, 
which engages with corresponding grooves in the bushing d, 
and thus rotates the bit on the back stroke. Fig. 159 shows 
the assemblage of parts of the rotation device, lettered as above. 
The valve is the same as described under the Ingersoll-Rand 
" Butterfly- Valve " drill (Chap. XX). An automatic oiler is 
shown at h. By the pulsations of the inlet air, oil is drawn 
into the valve chest, passing thence into the cylinder. Opera- 
tion of the drill is eased by the heavy springs /,/, connecting 
the front and back cylinder heads. The bit is held firmly in 
position in the chuck by pressure of the double spring g, but is 
readily changed. The handle has rubber grips. 

The Water Jackhamer '' BCRW-430" (Fig. 160) is Hke the 
above, except that it has a central water tube for delivering 
water under pressure to the hollow bit, as in the Leyner-Ingersoll 
drill, already described. It can drill holes to a depth of 10 or 
12 ft. The water tube is supphed with water through a swivel 
connection and a strainer. As the handle is offset, the tube is 



liAMMER DRILLS 



323 




Fig. 158.— IngersoU-Rand " Jackhamer " Hand Drill, '' BCR-430," for Dry Holes. 
6 




Fig. 159.— Rotation Device "of " Jackhamer " Drill. 



\W//////////A 




Fig. 160.— Ingersoll-Rand " Water Jackhamer," for Wet Holes. 
(Cylinder and accessory parts not shown.) 



324 



COMPRESSED AIR PLANT 



readily removable. A standard '' BCR " machine can be con- 
verted into a '' BCRW " water drill by substituting a few 
parts. The water pressure should be at least 25 or 30 lbs., 
but must always be less than the air pressure. Water tanks 
of 6 or 18 gals, capacity are furnished by the makers. The 
bit shank of the hollow steel for the Jackhamer drills is shown 




Fig. 161. — ''Jackhamer" Drill on Cradle Mounting. 



in the second cut of Fig. 153. Both of the above machines 
weigh 41 lbs. Diameter of steel, | in.; air hose, j in.; water 
hose, h in. 

For fiat-hole work, like drifting or breast-stoping, the Jack- 
hamer may be mounted as shown by Fig. 161. The cradle, 




Fig. 162. — " Jackhamer " on Special Mounting for Thin Coal Seams. 

somewhat resembling that of an ordinary rock-drill, is varied 
slightly in design according to whether it is to be used for the 
dry or wet machine; it is adapted to either column or tripod. 
A supplementary cradle, carrying the drill head, rides on the 
feed screw of the main cradle. Weight of mounting, 63 lbs. 

For mining thin coal seams the '' Jackhamer " drill is 
mounted as in Fig. 162. The legs of the mounting have adjust- 



HAMMER DRILLS 



325 



able extension ends, like those of a tripod. Cruciform steel, 
with a cross bit, and twisted as in Fig. 163, is used instead of 
ordinary octagon steel. When drilHng, the machine slides 
forward on the wooden frame, the bit being supported and 
held in alignment by a pair of guide clamps pivoted on the 
front end of the mounting. As the auger-like steel rotates, it 
assists in removing the cuttings from the hole. This new 
mode of mounting is well adapted to drilling the breast holes 
so common in coal mining. 



k- 



[<—m^^^\^2HM 



Length of Steel- 

E 



\<-2y^^. 




SECTION E-F 
Fig. 163. — Twisted Cruciform Bit, for Drilling in Coal. 

IngersoU-Rand '' Jackhamer Sinker " (Fig. 164) is a recent 
variation of the " Jackhamer," especially suited to shaft-sinking, 
or where deep holes (to 12 ft.) are required. It is built in two 
t3^es, dry and wet, both fitted with an axial tube for cleaning 
the hole with either air or water; in other respects they are 
identical in construction. In the dry machine, the air delivered 
to the tube is controlled by a valve in the handle. The water 
device is the same as in the Leyner-Ingersoll drill, and the 
rotation like that of the scandard " Jackharner." The valve 
works on the principle of the '' Butterfly," but is cylindrical, 
with end seats. As shown in the cut, the bit holder has a new 
form; to remove the bit, the yoke is swung to the right. 
Weight of drill, 70 lbs. ; size of steel, i in. 

IngersoU-Rand '' BuUmoose Jackhamer " is built especially 
for drilling deep down holes. Its design (Figs. 165 and 166) 



326 



COMPRESSED AIR PLANT 



is essentially different from the ordinary Jackhamer. A " But- 
terfly " valve is set at the back end of the drill, in line with the 
axis of the cylinder. The hammer reciprocates freely, the bit, 
which rests loosely in the chuck, being rotated independently, 




Fig. 164. "Jackhamer Sinker." Fig. 165. — " BuUmoose Jackhamer." 



as shown in Fig. 166. A small spool- valve a is operated through 
auxihary ports which are controlled by the movements of the 
hammer, and, by another set of ports, valve a throws the 
plunger b. A ratchet c, with four pawls, encircles the end of the 
bit chuck d, the ratchet in turn being inclosed by a steel ring e. 



HAMMER DRILLS 



327 



As the lug/, on ring e, engages with the plunger h,e h oscillated 
back and forth through a small arc, and by means of the ratchet 
thus rotates the bit. 

This drill has automatic lubrication and a blowing device 
for removing sludge from the hole. One-inch hollow steel is 
generally used. Weight of drill, 105 lbs. 




n Q 




Fig. 166. — " BuUmoose Jackhamer." 

Ingersoll-Rand " Imperial " Drills (types MV-i and MV-2) 
are valveless (Fig. 167). They resemble in principle the Hard- 
socg drill, alr.eady described, in that the piston (or hammer) 
acts as its own valve. The hollow piston a has an enlarged 
part h near the rear end, against the shoulder of which the air 
pressure is constantly acting, and a series of 6 slot-shaped ports 
c near the forward end. In the cut the piston has completed its 
stroke; the air is being exhausted through the piston ports c 
to the exhaust port d. The return stroke is caused by the 
constant air pressure on the shoulder h of the piston. 

The drill is rotated by a straight handle attached at e. The 
shank of the bit is held in a bushing /, which is made to receive 
either hexagonal or cruciform steel. Solid steel is used for these 
machines, which are designed for shallow '' plug holes," for 
mining and quarry service. Weight of drill, 42 lbs. 



328 



COMPRESSED AIR PLANT 



For drilling | to i-in. holes to a depth of 6 ins., as for block- 
holing, pop-shots, dressing walls of shafts, cutting timber hitches, 
etc., the Ingersoll-Rand Co. makes a small '' plug drill " (the 
" Invincible ")? weighing 21^ lbs. It is valveless, like the 
''Imperial." Downward pressure on the handle opens the 
throttle; the air is automatically shut oil when the drill is 
raised from the hole. The steel is rotated by the handle or by 
a wrench. The exhaust is led through a hose to the mouth of 
the hole, for removing the cuttings. 




Fig. 167.— Ingersoll-Rand "Imperial" Hammer Drill (Tj-pes MV-i, MV-2). 



Sullivan " Rotator " Hammer Drills are made in two forms, 
i.e., with air-tube or water- tube. 

The " Air-Tube Rotator " (Fig. 168) has a spool valve a, 
in a chest forming part of the cyhnder casting. The rotation 
device is similar to that of the Sullivan " DR-6 " drill (see 
description accompanying Fig. 154); ratchet b (Fig. 168) 
encircles the hammer c, which, by means of both straight and 
spiral grooving h and i, rotates the chuck bushing d. Air enters 
to the throttle at e, part of it going by the tube / to a passage 
in the rear cylinder head and thence through the axial air-tube 
g to the hollow bit. The bit is held in the chuck by the spring 
yoke j. Under the handle is an automatic lubricator k, like 
that used for the SulUvan " Liteweight " and '' Hy speed " drills. 
With each pulsation of air in the rear end of the cylinder, one 
of the small bails admits air to the oil chamber /, and the other 



HAMMER DRILLS 



329 



ball discharges a little oil into the cylinder, in the form of spray. 
Weight of drill, 7,S lbs. ; steel, | in. 

The " Water-Tube Rotator " (Fig. 169) resembles the 
" Air- Tube " machine so closely that a description is unneces- 
sary. It is intended especially for rather deep holes pointed 




Fig. 168.— Sullivan " Air-Tube Rotator " (DP-33) 

downward. Water, supplied by a tank under air pressure, 
like that in Fig. 136, Chap. XX, is admitted through the rear 
cylinder head to the axial tube. An inlet screen keeps out dirt 
and grit. A jet of live air mingles with the water, to increase 
the cleaning action. Weight of drill, 40 lbs.; steel, | in. 

Sullivan "Auger Rotator" (Fig. 170) is similar to the 
" PP-33 " drill, but has a shorter, Hghter and faster stroke. It 



330 



COMPRESSED AIR PL.\XT 



is designed for drilling in soft or broken rock (shales, coal, etc.). 
Solid, spiral steel is used, \\ath a forked (" fish-tail ") bit. It 
vnll drill 6-8-ft. holes in soft ore or coal, and up to 12 ft. in loose 
material. The piston is solid, no air or water tube being 
required, as the rotation of the spiral steel removes the cuttings 
from the hole. Weight of drill, 39 lbs. 




Fig. 169. — Sullivan "Water-Tube Rotator" (DP-33). 



Mountings for Sullivan Hammer Drills. Fig. 171 shows the 
" Rotator," with handle removed, mounted on a cradle attached 
to a column arm. The special cradle has two clamps; one [a) 
fitting over the cylinder, the other [b) bearing on the socket of 
the drill handle. A coil spring (c) at the rear of the shell acts 
as a shock absorber. Weight of cradle, 60 lbs. 



HAMMER DRILLS 



331 



Another form of mounting, the " pneumatic feed " (Fig. 172), 
consists of a light column, on the arm-saddle (d) of which is 
clamped a trunnion (e), with a hinged cradle-clamp, supporting 
a long air-feed cylinder (/). On the forward end of the piston 
rod of (/) is a short arm (g) and a saddle to which the drill is 




Fig. 170.— Sullivan " Auger Rotator," Class DR-33. 



clamped. The action of both drill and feed is controlled by a 
single throttle. 

''Hummer" Drills (Chicago Pneumatic Tool Co.). These 
are hand machines, with automatic rotation. Fig. 173 shows 
longitudinal sections of type " A-86," at right angles to each 
other. 



332 



COMPRESSED AIR PLANT 



Valve mechanism (Fig. 174). The hardened and ground 
steel ball B, weighing approximately i oz., reciprocates ^ in. 
in a steel cage F, provided with end seats. 5 is the inlet port to 
the valve chamber, C and P are the cylinder and hammer, Si 
and 52 the cylinder inlet ports, and £1, E2 the exhaust ports. 

Compressed air enters V through a series of peripheral holes 
Hj as shown by the arrows. The valve will thus be thrown 




Fig. 171. — Cradle and Column Mounting for Sullivan " Rotator " Drill. 



either forward or backward. Assuming that it takes its initial 
position as in the diagram, air flows through S2 to the cyHnder, 
and the hammer P makes its backward stroke. This movement 
of P uncovers exhaust port E2 and covers Ei. When air is 
exhausted through £2, an unbalanced condition is produced 
in the valve cage F, which causes valve B to move to its right- 
hand seat. Air from S then passes through S\ into the cylinder, 
and the hammer makes its forward stroke, thus completing 
the cycle of operation. 



HAMMER DRILLS 



333 



Rotation of the bit is independent of the hammer^ Refemng 
to Fig. 173, a small, high-speed, rotary air motor a is set trans^ 




Fig. 172. — Pneumatic-feed Mounting for Sullivan " Rotators." 

versely across the back cylinder head. The worm thread b, 
on the rotor shaft c, engages with a worm gear on the longitudinal 



334 



COMPRESSED MR PL.\NT 





/ 


[^ 






1 i 




o 
U 

o 

H 

o 
'■Xj 



O 

to 
a 
u 

IS 

U 



HAMMER DRILLS 



335 



shaft d. At the other end of tf is a pinion e, meshing with a gear 
mounted on and encirding the chuck bushing /, which holds the 
hexagonal shank of the bit. There is no ratchet. The air 
admitted to the drill goes first to the rotation motor; the exhaust 
from this passes through S (Fig. 174) to the valve B, for operating 
the hammer. The bit holder g (Fig. 173) is similar in general 
design to that of several of the machines previously described. 

For drilHng in coal, slate, or other soft rock, a twist or auger 
bit may be used (see Ingersoll-Rand '' Jackhamer " and Sullivan 
"Auger Rotator ")• 

The Chicago Pneumatic Tool Co. makes several other types 
of hammer drill. 



Front Enct 




El E2 

Fig. i'74. — Valve Mechanism of "Hummer" Drill (Diagrammatic). 



Waugh Hammer Drills (Denver Rock Drill Manufacturing 
Co.) are made of several types. (Waugh " S topers," with tele- 
scopic air-feed standards, are described later.) The " Clipper " 
drill, unmounted, weighs 47 lbs.; the " Dreadnaught," 2>2, lbs. 
Their general design is essentially the same. As hand drills, 
they serve for shaft-sinking, blockhohng, quarrying, etc.; 
mounted on a guide shell (on column or tripod) , they are appHc- 
able to horizontal breast work and drifting. 

The " Clipper " drill. Model 50 (Fig. 175), is valveless. Air 
enters at a. The hollow hammer h has four ports, as shown. 
In making the forward stroke, air acts on the entire area of the 
hammer; for the back stroke, it acts on the shoulder c. Rotation 



336 



COMPRESSED AIR PLANT 




o 

to 



-^3 
O 



a. 
.9r 
U 






O 



HAMMER DRILLS 337 

is effected by the ratchet d (in the front cyHnder head), which 
encircles the rifled extension e of the hammer. As the hammer is 
prevented from rotating by two longitudinal splines in the 
cylinder, each stroke of the hammer turns the rifle-nut/, which 
contains the pawls and their springs. In the forward edge 
of the ratchet ring ^ is a set of small lugs, which engage with 
lugs on the chuck bushing Ji, holding the hexagonal bit shank. 

The water tube i receives water at 7, the supply being con- 
trolled by the needle valve g (see detail cut). As the tube is 
reduced in size at the point k, making a loose fit thence to the 
end of the hammer, air mixes with the cleaning water. To 
increase the quantity of water, when required, the drill is raised 
slightly, allowing the hammer to run forward and thus uncover 
more of the smaller part of the tube i. Model 55 of the Waugh 
drill has no water attachment, a blow-valve being used instead, 
to deliver a large quantity of air at intervals, when drilling 
deep holes. 

The '' Dreadnaught " hand-hammer drill (Model 60) is simi- 
lar in general design to the '' Clipper," but is larger and heavier. 
Another type of "Dreadnaught," having a valve and an air-feed 
standard, is described on p. 340. 

McKiernan-Terry *'F-I" Hammer Drill (Fig. 176) has 
automatic rotation, uses hollow, J-in. steel, and will drill to a 
depth of 8 or 10 ft. Weight, ^'^ lbs. The spool valve h is small 
and light, with a throw of only \ in. The throttle a has three 
positions: the first shuts off air; the second opens the port c, 
to admit air to the hollow bit, for cleaning the hole; the third 
starts the drill. When blowing out the drill hole, the bit is 
shghtly raised from the bottom. From port c the air for cleaning 
goes to the forward end of the cylinder through passage d, 
alongside of the hammer. Rotation is similar to that of the 
Leyner-Ingersoll drill (Fig. 152) ; e is the rifle-bar,/ the rifle-nut, 
in the hammer g; the front end of g has longitudinal fluting h, 
which engages with corresponding grooves in the chuck bushing 
or socket i. This bushing receives the hexagonal shank of the 
bit. To remove the bit, the yoke 7, normally held in position by 
the springs k, is swung to the right. 



338 



COMPRESSED AIR PLANT 



McKiernan-Terry '' A-g " Hammer Drill (Fig. 177), weigh- 
ing 90 lbs., is designed for drilling deeper and larger holes than 
the preceding. The spool-valve is h'ke that of many of the 
standard reciprocating drills. Aside from this, though there are 




HUJ 



ucr 



P'iG. 176. — McKiernan-Terry " F-i " Hammer Drill. 



differences in the details of construction, the main features are 
the same as in the " F-i " drill. Size of steel used, i in. The 
steel is not specially shanked, the rotating chuck bushing receiv- 
ing the full hexagonal section. 

Another drill by the same makers is the " Busy Bee " f '' B-i ") 



HAMMER DRILLS 



3S9 



weighing 50 lbs. It is designed for shallower holes, and for 
dressing or trimming, blockholing and similar work. 

Wood Hammer Drill has a spool- valve, of essentially the 
same design as the Wood reciprocating drill (Chap. XX). In 
its general features, including the mode of rotating the bit, 
it is quite similar to the McKiernan- 
Terry " A-9 " drill, described above. 
Weight, 44 lbs.; diameter of cylinder, 2 ins. 

Class C. Stopers, or Hammer Drills 
WITH Air-Feed Standards 



General Description. These machines 
are made by most of the rock-drill manu- 
facturers. They are intended chiefly for 
drilling holes directed above the hori- 
zontal, as in overhand s toping, though 
they may also be mounted on a column, 
for drifting and other breast drilling. 

Nearly all makes are built on the same 
general lines. The design of the drill itself 
is usually the same as that of the hand- 
hammer drills of the same maker. At- 
tached to the back head of the drill is a 
long, telescopic extension, on the end of 
which the machine stands when drilHng 
overhead holes. This standard comprises 
a slender cylinder and piston, which when 
supplied with compressed air automati- 
cally keep the drill fed up to its work. Fig. 177. — McKieman- 

Air for both drill and feed cylinder is ^^''^ "^'9" Ham- 

. .11 1 1 i^^i" Drill- 

admitted by a single throttle valve. 

Most of these drills have no automatic rotation; to keep 

the hole round, the entire machine is rotated on its axis by an 

arm or handle. They are all one-man machines. A few makes 

use a water attachment, for directing a spray of water at the 

mouth of the hole, to moisten the dust.* Others use hollow 

* In this connection, see the " dust allayer " of the Climax drill, p. 280. 




340 COMPRESSED AIR PLANT 

steel, with a water tube passing through the axis of the drill, 
like many of the drills in Classes A and B . 

Owing to the similarity in design of the air-feed of the dif- 
ferent makes, and to avoid repetition, the reader is referred 
to the following cuts for details of construction. The descriptions 
of typical drills given below are confined chiefly to the valve 
motion and the operation of the drill itself. 

Waugh Stoper is made in several sizes and weights. The 
lightest model (2|-in. cyHnder), using i-in. steel and striking a 
short, rapid blow, is suitable for the softer rocks and ore. The 
heaviest machine (2f-in.) has longer stroke, uses i| or ij-in. 
steel, and strikes a heavier blow, as needed for hard ground. 

A recent model of the Waugh stoper is the No. 14A (2j-in.). 
Its operation is shown by Figs. 178, 179 and 180. The valve A 
is a hollow cylinder, with external annular recesses and a rear 
collar r. On the back stroke, the neck w of the hammer enters 
the bore of the hollow valve. Compressed air is admitted 
to the drill by the path shown by the arrows, passing through 
the channels a, b, c and d, to the valve, which then takes the 
rearward position, as in Fig. 178. This position is caused by 
a differential pressure on the valve, due to the small difference 
in the diameters and areas of the portions y and z. The live 
air from the recess or chamber c is then free to flow into the 
rear end of the cyHnder c (Fig. 179), since the diameter of 
the hammer neck w is about -^ in. less than the diameter x 
of the valve chest. As the hammer is driven forward, any 
air in the forward end of the cylinder remaining from the pre- 
ceding stroke is exhausted through ports and passages g, h, i, j 
(Fig. 180), annular groove k of the valve (Fig. 178), and exhaust 
ports / and m (Fig. 179), to the atmosphere. 

When the hammer in its forward stroke uncovers the opening 
^(the '' trip-hole ") (Fig. 180), air flows through passages n, 0, 
and p, into chamber q, where it acts on the large end r of the 
valve and shifts it into its forward position. Air then passes 
from b, through d and recess k (Fig. 178), and thence through 
7, i, h and g (Fig. j8o) to the front end / of the cylinder, thus 
driving the hammer back. During this stroke, the air in the 



HAMMER DRILLS 



341 




r 



il ! 



--fH: 



l<^' <.^^ 





■o 
V 

be 



P 



o 
> 



< 
^ 



O 



0) 

a 

o 

• -I-" 

en 

be 

::! 

c3 



00 

H 

6 

M 



342 



COMPRESSED AIR PLANT 



rear end of the cylinder is exhausted through the hollow valve 
into chamber q (Fig. 179), and thence out of the exhaust ports 
/ and m. As the exhaust is not yet down to atmospheric pres- 
sure, it holds the valve in its forward position, but, when the 
hammer neck w enters the valve, the exhaust pressure in q drops 
to atmospheric pressure, and no longer holds the valve. There- 




rig. 179 



Note:- These Sections are Diagrrammaticj 
Hammer not shown. There are 
actually : 2 Ports g and h ; 
6 Ports d and b ; 
l.Pocts ni and 1 Ports I, 




Eiff. 180 o 



Figs. 179 and i8o. — Waugh Stoper, Model 14A.. Diagrammatic Sections of 

Cylinder, Valve and Ports. 



upon, the differential pressure on areas y and z (Fig. 178) be- 
comes effective, and the valve shifts to its rearward position. 
Air then enters the rear end of the cylinder at e, and the hammer 
is again driven forward, thus completing the cycle. Near the 
end ot both forward and back strokes air exhausts expansively 
through ports s, t and m (Fig. 179). The small passage w is a 
vent for the chamber q. 



HAMMER DRILLS 



343 



Ingersoll-Rand Stope Drills are of two types, the chief 
f erence being in the design 
of the cylinder. The air 
feed, made in three forms, 
is the same for both, as 
noted below. 

The "Butterfly" 
stoper (Fig. i8i) is of 
solid bolted construction, 
except the front head, 
which is spring retained. 
An " anvil block " a is 
interposed between the 
hammer b and the drill 
bit c. Air is admitted 
from the inlet d by the 
throttle e to the " Butter- 
fly " valve /, the ports 
being shown by the dotted 
areas opposite each wing 
of the valve. One wing 
controls admission to both 
ends of the cylinder, the 
other controlling the ex- 
haust. Details of the 
operation of this valve 
are given in Figs. 141- 
143 (Chap. XX), with 
accompanying descrip- 
tions. The exhaust open- 
ing g directs the exhaust 
backward, to minimize its 
tendency to scatter the 
dust at the mouth of the 
drill hole. Throttle e is 
so designed that, when in 
position to admit air to 



dif- 




(U 

(X 
o 

C/2 






3 

m 






(U 



o 

t-( 



344 



COMPRESSED AIR PLANT 



the drill (as in Fig. i8i), the small port h is also opened, to allow 
air to pass to the feed cylinder. 

The rotating handle, shown in cross-section and detail 
drawing at i (Fig. i8i), and in longitudinal section by Fig. 182, 
contains an automatic oiling device, operating as follows: In 
the bore of the handle is a porous plug 7, which regulates the 
flow of oil and strains out dirt and grit. The oil chamber /, 
filled once a shift by removing the screw plug w, is in communi- 
cation by the small port k with the live air side of the valve 
chest. When the drill is at work, the air pulsations in the 
chest draw the oil through j and the small passages n and k 
into the chest, whence it passes with the air into the drill cylinder. 
In the valve chest, close to the throttle, is an air strainer 



Inlet 




Fig. 182. — Oiling Device, " Butterfly " Stoper. 



(see detail cut in Fig. 181), consisting of a perforated cup- 
shaped disk 0, held in position by the spring p.. 

Three forms of air feed are made: (i) the cylinder feeds off 
the piston (Fig. 181), which allows the drill to be mounted on 
a column or tripod, for making breast holes; (2) the piston feeds 
out of the cylinder; (3) an extension point is added, for drilling 
uppers in workings with a high roof. Referring to Fig. 181, 
the hollow piston g is a steel tube, with a flange for bolting to the 
drill body. The stuffing box r of the cylinder s contains two 
cup-leathers. When the piston and cylinder close up, they are 
automatically held together by the friction spring / entering 
the recess u. Fig. 183 shows the air feed of the second type. 

The " Butterfly " stoper uses i, i| or ij-in. cruciform steel 
(generally solid), with square, hexagonal or cruciform shank. 
Weight of drill, 74 lbs. 



HAMMER DRILLS 



345 




Fig. 183.— Air Feed for " Butterfly " 
Stoper, Type " BC-21." 







Fig. 185.— Cylinder and Valve of 
" CC Stopehamer." 






B 
o 

u 









00 



O 

M 

fin 



346 



COMPRESSED AIR PLAXT 



The Ingersoll-Rand "* CC Stopehamer "" (Fig. 184) is similar 
to the " Butterfly " stoper. except in the design of cyhnder and 
valve. The cyhnder is a drop forging; a is the haninier. b the 
" amil block.'" c the inlet elbow. The valve d has a very short 
throw and is designed to give a high degi'ee of expansion in the 
use of the ah-; it has end seats. Fig. 185 shows the cylinder, 
valve and ports in more detail. 

A wet t}-pe of ''' Stopehamer " fCCW) is also made, the 
water de^^ce being the same as in the " Butterfly "' stoper. 
Fig. 186 shows a spray dust-allayer, which may be attached to 




'\ 


Spra- — .- 


»"=(^ 


1 


^-J'N 





Air Supply 



Water Hose 




Fig. 186. — Dust Allayer for '' Stopehamer." 



the '' CC " drill by replacing the inlet elbow by a special con- 
nection. Weight of the '' Stopehamers "' is 81 lbs. 

''Chicago Stoper" (Chicago Pneumatic Tool Co.), with 
air-feed standard, has the same valve motion as the '' Catling " 
drill of the same makers (Chap. XX. Fig. 144). As it resembles 
in its general features the drills of this class described on preced- 
ing pages, details are omitted. It ranks \\ith the best of the 
stoping drills. Sohd, cruciform steel is used, and a spray dust- 
allayer is pro\'ided, which is readily attached when desired. 
Weight of drill. 70 lbs.; diameter of cylinder, 2| ins. 

Sullivan Hammer Drills with air-feed. In these the feed 
cylinder is not an integral part of the drill. For details and 
illustrations, see Sullivan drills under Class B. 

Cochise Air-Feed Drill (Cochise Machine Co., Los Angeles, 
Cal.) is similar in general design to the other typical stopers 
with air-feed standards, as described on preceding pages. It 



HAMMER DRILLS ' 347 

has a spool valve. Weight of drill, 73 lbs.; diameter of cylinder, 
2J ins. 

Operation of Hammer Drills 

Air Consumption of Hammer Drills is approximately the 
same as for reciprocating drills of the same cylinder diameter 
(see Tables XXXI and XXXII, Chap. XX). But, in comparing 
the results in terms of work done, it must be remembered that 
the smaller sizes of Class A drills, and all of Classes B and C, are 
one-man machines, and that for some kinds of work, especially 
overhead stoping, the hammer drills make a higher footage 
of hole. 

Depth of Hole. When hammer drills were first introduced, 
it was found that the speed of drilling materially decreased at 
depths greater than 3 or 4 ft., even with the use of a water jet 
alongside of the bit. This was due in part to the inertia of long 
and heavy bits, but probably more to the failure of the early 
drills to " mud " well. In recent years the designs have been so 
improved that, even for " down " holes, they can now be used 
for depths of at least 10 ft., though average depths are usually 
from 4-6 ft. This is largely the result of using hollow steel 
and a strong jet of water mixed with compressed air. The force 
of the expanding air assists in keeping the hole clear of pasty 
sludge, thus allowing the hammer to strike a more effective 
blow. As a rule, the fastest work is done when drilling " uppers " 
(holes directed at a steep upward angle) in dry rock or ore. 
The dust and cuttings then run out by gravity; that is, the 
holes are self- cleaning. 

Records of Work. The prefatory remarks, made under this 
heading in Chap. XX, respecting reciprocating drills, apply 
also here. But, since hammer drills are usually operated with- 
out mounting, no allowance of time for setting up is necessary; 
and, as there are no chuck bolts to manipulate, bits can be 
changed in i|-2 minutes. Table XXXVIII gives figures based 
on a number of recorded runs. Type of drill (column 3) refers 
to the classification used in this chapter. 



348 



COMPRESSED AIR PLANT 



Table XXXVIII 



Kind of Work. 



Stoping 

Same as above 

Drifting 

Drifting 

Stoping 

Stoping , 

Stoping 

Stoping 

Drifting 

Stoping 

Drifting 

Drifting 

Stoping 

Stoping 

Stoping 

Tunnelling 
(test run 
Sinking. . . 
Sinking . . . 
Drifting. . . 
Raising . . . 



Rock or Ore. 



Hard phonolite breccia ... 
but unfavorable conditions. 
Amygdaloid copper rock. . . 

Tough schist 

Hard trachyte . . 

Hard trachyte 

Hard porphyry 

Am^-gdaloid copper rock.. . 
Amygdaloid copper rock. . . 

Quartzose ore 

Granite 

Quartzose ore 

Medium andesite 

Pyritic ore, medium hard . . 
Pyritic ore, medium hard . . 



Average granite 

Tough schist . . . 
Tough schist . . . 
Hard limestone. 
Quartzose ore . . 



Type of 
Drill. 



Class C 



Class A 
Class A 
Class C 
Class C 
Class C 
Class C 
Class A 
Class C 
Class A 
Class A 
Class C 
Class C 
Class C 

Class A 

Class B 
Class A 
Class A 
Class C 



Depth of 

Hole, 

Ft. 



Inches of Hole 
per Minute. 



2. 2 
2.6 
6.0 



4.0 
50 
S-5 
6.0 
6.0 
30 
30 
3-3 
4-3 
5-3 
50 



6.5 
7-7 
3-2 
4.0 



Total 
Time. 



2.04 
1.08 

0. 7-2 

2.64 

2-35 

1 . 20 
2.09 
0.93 
215 
215 
3.00 

1-58 
2.85 
2.00 
1 . 10 

3 56 

1-43 
1 .00 

163 
2.40 



Net 
Time. 



2.54 



2.90 



3.22 
3 58 



3-52 



=;.i6 



1.86 
5 10 



Field of Work. For driving tunnels, drifts and crosscuts, 
or for underhand stoping in wide veins, and wherever deep 
holes of large diameter can be advantageously adopted, recipro- 
cating drills are still in general use, though the same field of 
work is occupied by the hammer drills of Class A. In connection 
with these operations, hammer drills are useful as auxiliaries, 
for blockholing and for '' squaring up " after the main rounds 
have been fired; that is, dressing the walls and taking up the 
*' bottom " when the deep holes fail to break clean. 

Most of the hammer drills, particularly those of Classes 
B and C, are less well adapted to making holes that approach 
the horizontal, as in tunnelling, drifting, crosscutting and breast 
stoping. Class B machines are best for down holes, as in 
quarrying and trenching in rock. They are also used success- 



HAMMER DRILLS 349 

fully for shaft sinking in rather soft and laminated or thin- 
bedded rock (like shales), or where there are many slips and 
short fissures. In such rocks, a relatively large number of 
shallow holes give the best results, and to drill them with 
reciprocating machines involves extra loss of time for 3hifting 
and setting up. 

Class C drills are especially designed for, and do their best 
work in, drilling uppers, as in general overhand stoping. They 
are useful, also, in mining thin veins with narrow pay streaks. 
For breast holes, they are sometimes mounted on columns. 

In drilling dry holes, hammer drills raise much dust. With 
the hand machine, the operator must stand close to the mouth 
of the hole, and, when hollow steel is used with an air jet, the 
dust is blown back in the operator's face. This is especially 
troublesome in drilling uppers with Class C machines. The 
harmful effects of rock dust containing siliceous matter, also 
the spray dust-allayers and other water devices of hammer 
drills, have been referred to on preceding pages. 

Solid bits have a limited application. They can be used for 
shallow, holes in the softer rocks, by watering the hole and 
spooning out the sludge; or in dry rock, provided the holes are 
at a sufficient upward angle to permit the cuttings to run out by 
gravity. But they are best used for cutting hitches for mine 
timbers, blockholing, and quarry work. 

Makers of Hammer Drills. The following alphabetical list 
includes all of the principal American makers: 

C. T. Carnahan Mfg. Co., Denver, Colo. Her Rock Drill Mfg. Co., Denver, Colo. 

Chicago Pneumatic Tool Co., Chicago, 111. Ingersoll-Rand Co., New York 

Cleveland Pneumatic Tool Co., Cleveland, O. McKiernan- Terry Drill Co., New York 

Cochise Machine Co., Los Angeles, Cal. Shaw Pneumatic Tool Co., Denver, Colo. 
Denver Rock Drill and Mach. Co., Denver, Colo. Sullivan Machinery Co., Chicago, 111. 

Flottman & Co., Cardiff, Wales R. Stephens & Son, Carn Brea, Corn- 
Hardsocg Wonder Drill Co., Ottumwa, Iowa wall, England 

Whitcomb Hammer Drill Co., Rochelle, 111. 



CHAPTER XXII 
COAL-CUTTING MACHINERY 

Coal-cutting machines have largely replaced hand labor 
in " under-cutting " the coal, preparatory to breaking it by 
blasting or wedging. Objects: (i) To economize in the cost of 
mining; (2) to decrease the proportion of " fines " produced; 
(3) to increase the rate of production from a given extent of 
mine workings. Coal cutters are used chiefly in bituminous 
collieries, and, unless wages are very low, they can mine more 
cheaply than by manual labor. In recent years they have also 
been successfully employed, to a limited extent, for mining 
anthracite in veins of rather flat pitch. They groove or 
undercut the face or breast of coal, close to the floor and to a 
depth of several feet. The mass so undercut is subsequently 
broken down by comparatively light blasts. 

Coal cutters comprise four classes: (i) Endless-chain; 
(2) Rotary-bar; (3) Disk; (4) Reciprocating or pick machines. 
Machines of the first three classes are driven by electricity or 
compressed air, electricity being now in most general use. Pick 
machines are operated by compressed air, no satisfactory electric- 
driven pick having yet been brought out.* 

Endless-Chain Cutters are built by several makers in the 
United States, among whom are the: Goodman Manufacturing 
Co., Jeffrey Manufacturing Co., Morgan- Gardner Electric Co., 
and Sullivan Machinery Co. Most of the machines of this type 
are electric-driven, though the Jeffrey and Sullivan chain cutters 
are furnished with compressed-air drive for use in gaseous mines, 
or where local conditions make it convenient. 

* The " Pneumelectric Coal Puncher " (see p. 372) is not a true compressed- 
air pick machine, in the sense in which the term is here used. 

350 



COAL-CUTTING MACHINERY 



351 



The standard chain machines now in use are self-propelling 
along the face of coal to be undercut. Differences in design 
are chiefly matters of detail, rather than of principle. For 
shifting and feeding the machines, there are small wire-rope 
or chain drums, mounted on the rear end and operated by 
gearing from the air engine or electric motor. The rope or 
chain is made fast to a timber, or held by a '' jack " set between 



Coal 




ROOM-AND-PILLAR MINING WITH 
OLD STYLE BREAST MACHINE 



ROOM-AND-PILLAR MINING WITH 
SULLIVAN CONTINUOUS COAL CUTTER 



Fig. 187. 



roof and floor, at any desired distance from the machine. By 
throwing the drum into gear, the machine is thus pulled (fed) 
along the face, as the undercut advances. 

Some machines have but one drum; others (as the Jeffrey 
*' Longwall ") have a special " sumping " drum, by which the 
machine is brought into position at the point where cutting is to 
be started, and is then fed forward or '' sumped " into the face S(, 
distance equal to the depth of the slice to be taken. 



352 



COMPRESSED AIR PLANT 



Figs. 187, 188 and 189 show different modes of manipulating 
a chain machine in room work. 

The makers of coal cutters furnish special trucks for moving 
the machines from place to place underground. Some of the 
trucks are propelled by a sprocket chain, driven from the cutter 
engine (see Fig. 193). 

The Jeffrey " Longwall " chain machine, for either electric 

or compressed-air drive, is especially 
designed for continuous undercutting 
on long faces, as in longwall mining. 
Figs. 190, 191, and 192 show the 
compressed-air driven type.* 

the front end of the machine 
92) is pivoted the cutter frame 




position of sullivan longwall ironclad in cutting along the face 

Fig. 188. 

a, carrying the chain. The chain runs on two sprockets, 
one at each end of the cutter frame (Fig. 190). At the 
inner end is the driving sprocket, operated through worm 
gearing by the compressed-air turbine // (Fig. 192). By means 
of the jaw clutch /, the cutter chain can be started or stopped, 
as required. The turbine (Fig. 191) consists of two rotors, with 
helical blades or teeth cast on their surfaces, and is designed to 
work with air pressures of 40-80 lbs. 

* At the present time (1918), due to war conditions, the Jeffrey Co. is not 
building air-driven machines. 



COAL-CUTTING MACHINERY 



353 




354 



COMPRESSED AIR PLAXT 



I 



/ 



o 
U 






1 Wl' 



.^^ 



**r 



I 



i-^^ 



.\' 



7^X. 



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COAL-CUTTING MACHINERY 



355 



Another worm i (Fig. 192), on the opposite end of the turbine 
shaft g, operates the feed and sumping drums 7 and k, as follows. 
The feed worm-wheel / has a crank pin m and connecting rod 




Fig. 191. — Compressed- Air Turbine for Longwall Coal Cutter, Jeffrey Mfg. Co. 

w, which is pinned to the ratchet drive-lever 0, carrying a pawl 
p engaging with the ratchet q. On the lower end of the vertical 
spindle r of this ratchet is a pinion s, engaging with a gear t, 




SECTION A-A 



Fig. 192. — ^Diagram of Operating Mechanism, Jeffrey Coal Cutter, Type 24-A. 

which is attached to the cone clutch v. The hand wheel u 
operates the drums by forcing them down into contact with 
the cone clutch. 



356 



COMPRESSED AIR PLANT 



This machine is made in two sizes. The 24-A size is 8 ft. 
2 ins. long, 31 ins. wdde, 19 ins. high; weight, 6,000 lbs.; lengths 
of cutter arm, 4, 5, and 6 ft. The 36-A size, especially designed 




I 

CO 



3 

u 

o 
U 



o 

C/3 









o 

t— I 



for thin, pitching seams, is 8 ft. 8 ins. long, 31 ins. wide, 15 ins. 
high; weight, 4,000 lbs.; lengths of cutter arm, 3, 4 and 5 ft. 

The Jeffrey " Shortwall " Coal Cutter, type 35-B (Fig. 193), 
is built for both air and electric drive, and is designed especially 



COAL-CUTTING MACHINERY 357 

for room work. Fig. 189 shows the different modes of starting 
a cut. The cutter head is rigidly attached to the frame, the 
entire machine being swung in following its work. The chain 
is driven through worm gearing by a compressed-air turbine, 
similar in general to that used for the longwall cutter. Fig. 
193 shows two views of this machine, as mounted on a self- 
propelling truck for delivery at the point where the work is to 
be done. 

Fig. 194 shows an older type of Jeffrey chain machine, for 
room or breast work, and still used in a number of mines in the 
United States. It has a bed frame of two parallel steel chan- 
nels, with cross braces, within which is a T-shaped sliding frame, 
carrying on the rear end a small duplex air engine. The sliding 
frame carries an endless sprocket-chain, driven by gearing from 
the engine. Sockets in alternate links of the chain hold the 
cutting teeth or bits. In the forward end of the sliding frame 
are two idler sprockets carrying the chain. On each end of the 
main frame are screw-jacks, for bracing the machine in position. 
The bits are so shaped and staggered as to " cover " the chain 
and cutter head, and make a groove in the coal about 4 ins. 
high, or sufficient to permit the cutter head to enter the undercut 
freely. The sliding frame is fed forward by a pair of pinions 
engaging with feed racks on the stationary frame, and driven by 
a worm gear from the engine. 

Depth of undercut, 4-7 ft., width, 39-44 ins. By shifting the 
machine, sidewise, successive cuts are made along the face. 
From 100-150 sq. yds. can be undercut in 10 hours. Power 
required, 8-14 H.P. 

The Sullivan " Ironclad " coal cutters are built for both 
electric and compressed-air drive. Fig. 195 shows the longwall 
machine, class CH-8, operated by an air turbine. The Sulli- 
van Machinery Co. also builds an air-driven chain machine for 
room-and-pillar work, class CE-7 (Fig. 195a). Figs. 187 and 
188 show diagrammatically these two types, operating respec- 
tively in room-and-pillar and longwall mining. 

In its general lines, the air turbine is similar to that used for 
the Jeffrey longwall machine, already described; it consists 



358 



COMPRESSED AIR PLANT 




COAL-CUTTING MACHINERY 



359 



essentially of a pair of wide helical gears, ground to each other 
and making a close fit in the casing. 

Besides the machines already described, several other types 
of chain cutters are made in the United States, but, as they are 







u 



3 
a 

o 
U 



be 
o 



T3 
CI 

o 









electric-driven, any discussion of them would be out of place 
in this book. 

Rotary-Bar Cutters. A design formerly used to a limited ex- 
tent in the United States for room work, but no longer made 



360 



COMPRESSED AIR PLANT 




COAL-CUTTING MACHINERY 



361 



here, had the same general lines as the chain machine in Fig. 194. 
The sliding cutter head carried a transverse bar, driven by a 
sprocket chain and in which teeth were set. Length of cutter 
bar, 7,-7,^ ft., speed of rotation, 200 revs, per min., depth of cut, 
4-5^ ft. 

Table XXXIX 
General Dimensions of Sullivan Air-Driven Coal Cutters 



Air pressure, lbs 

Horse-power of motor 

Cutter bar, length, ins 

Distance, face to props, ins. . . . 

Height when cutting, ins 

Height on standard truck, ins. 
Height on drop-axle truck, ins, 

Height of groo\,'e cut, ins 

Feed along face, ins. per min. . 
Weight, machine only, lbs ... . 

Weight, truck only, lbs 

Weight, equipment only, lbs . . 



Class CE-7. 
Room Work. 



20-70 

30 

66, 78 or Qo 

70 • 

24I 

34f 

3of 

5f 

15-39 

5800 

1000 

850 



Class CH-8. 
Longwall. 



20-70 

30 

24, 30, 36, 42, 

48,54,60 or 66 
30 

28I 
24! 

5 
12-36 
5300 

950 

850 



Fig. 196 shows a recent British bar cutter of a wholly different 
design, built for longwall mining by Mavor & Coulson, Glasgow, 
Scotland. It may be mounted on runners, sliding on the floor, 
or on wheels for travelling on a track. By mounting it on skids, 
as in the lower part of the illustration, the cut can be made in 
any part of the seam, between floor and roof. The machine 
will work in seams up to a pitch of about 60°. Fig. 197 shows 
the mode of operation. 

The cutter bar, which is driven through gearing by a double- 
cyHnder air engine, or an electric motor, can be swung through 
a total horizontal arc of somewhat more than 180°. The bar is 
threaded spirally, to remove the cuttings from the groove, and 
staggered bits are set in sockets throughout its length. 

This machine is made in 3 sizes, ranging from 7 ft. 9 ins. 
to II ft. long (without the cutter bar), i ft. 4 ins. to 2 ft. high, 



362 



COMPRESSED AIR PLANT 



and 3 ft. 4 ins. to 3 ft. 11 ins. wide; shipping weight, 3,640 to 
6,160 lbs.; maximum depth of undercut, 3^-6 ft.; air pressure 
at the machine, 45 lbs. 



^^i*^|L.-^3f- 



^l 



l-^'i 



i^HM 




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PQ 

I 

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On 






Disk Cutters are a very old type, invented in Great Britain. 
A machine formerly made by the Jeffrey Manufacturing Co. is 
shown in Fig. 198. Disk cutters are still employed to some 



COAL-CUTTING MACHINERY 



363 



extent in Europe, but are almost obsolete in the United States. 
They are designed for longwall mining only, and are especially 
useful in very thin, steeply pitching seams, being fed along the 




3 

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o 

o 
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a 



(V 






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C 



ID 
O 



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face by their own power, travelling back and forth, up and down 
the dip. 

The cutter disk is supported by a bracket on one side of the 
frame, and projects a distance nearly equal to its own diameter. 
On the periphery is a series of staggered bits, cutting a groove 



364 



COMPRESSED AIR PLANT 




COAL-CUTTING MACHINERY 



365 



high enough to admit the disk freely. The speed of the wheel 
is usually 15-30 revs, per min., depending on the character of 
the coal. When in operation the machine is automatically 
pulled along the face of coal by a wire rope, made fast to a post 
at the end of the face, and wound in by a drum geared to the 
driving engine. 

A recent British disk cutter, built by Mavor & Coulson, 
Glasgow, is driven by electricity only. 

Pick Machines in general construction resemble the recipro- 
cating rock drills; but all work without rotation of the bit, 




Fig. 199. — Sullivan Coal Pick, Working in a Thin Vein. 



since in undercutting there is no question of preventing rifling 
of the hole, as in rock-drilling. Fig. 199 shows a typical case of 
a pick machine in operation. The machine is mounted on a 
wooden platform, 3 ft. wide by 8 ft. long, which slopes towards 
the face of coal, at an angle of about 5°. The recoil of the 
blows is thus nearly neutralized by gravity and the machine 
kept up to its work The operator chocks the wheels with 
wooden blocks, sometimes strapped to his boots, and directs 
the blows by swinging the machine laterally, with the wheels 
as a fulcrum. To give the machine sufficient reach, the front 



366 ' COMI^RESSED AIR PLANT 

cylinder head and piston rod are very long. A horizontal width 
of 4 or 5 ft. of undercut is thus readily commanded. Depth of 
cut rarely exceeds 5 ft. A helper clears away the debris with a 
long-handle shovel, and assists in moving and setting up the 
machine. 

Most pick machines run at 200-250 strokes per min. The 
lower speed machines probably have some advantage, because, 
as each individual blow is directed by the operator, he can 
increase the efl&ciency of the work if he has time between strokes 
to point the pick in such a manner that it will do most execution. 
In average coal, an undercut of 4 by 4 ft. in horizontal area 
can be made in 16-18 mins. The platform can be shifted 
sidewise to the next position and the bit changed in 8-10 mins. 
Height of undercut is 12-14 ins. at the face, tapering to 3 or 3 J 
ins. at the bottom. Under favorable conditions, good operators 
can undercut, per shift, 75-85 linear ft. of face, to a depth of 
4-4I ft.; fair, average work, 60-65 ft. of undercut, 4 ft. deep. 

The Harrison Pick machine, one of the oldest of this class, 
was invented in 1877. It has been built for many years by the 
George D. Whitcomb Co., Rochelle, 111. Fig. 200 shows the 
longitudinal section. The valve is a long, double spool, actuated 
through a crank and connecting rod by a horizontal rotary 
engine set above the valve chest. The main cylinder of the 
machine has double ports at each end to cushion the stroke, 
and for running with a short stroke when desired. 

These machines are made in heavy and light patterns, 
weighing respectively about 700 and 500 lbs. The heavier pick 
will cut to a depth of 5 ft. and is especially adapted to " shear- 
ing "; that is, making a vertical cut or groove, on one or both 
sides, in driving entries. For this purpose, the supporting wheels 
are 34 or 40 ins. diameter, to raise the machine high enough to 
give the requisite reach. Smaller wheels are used for ordinary 
undercutting. 

The Sullivan Pick (Fig. 201) is made in 3 sizes, with cylinder 
bores of 4I, 4I and 5I ins., for undercutting to maximum depths 
of 4^-6 ft. Weights, 650-850 lbs. Air consumption: 4|-in. 
machine, about no cu.ft. free air per min.; 5|-in. machine, 



COAL-CUTTING MACHINERY 



367 



130 cu.ft. Standard 
wheels, 12-24 iiis. diam- 
eter; wheels for shear- 
ing, 26-48 ins. Fig. 199 
shows the pick at work. 
The spool-valve (123) 
throws the long flat 
valve (126), which con- 
trols the main ports 
(148, 149) and the sec- 
ondary ports between 
them. These secondary 
ports produce cushion- 
ing at each end of the 
stroke, as follows: A 
check-valve (130) is in- 
serted in the forward 
main port (149), and, 
when the pick does not 
strike the coal, the 
piston runs forward far 
enough to form an air 
cushion in the front end 
of the cylinder. This 
closes the check-valve, 
and prevents immediate 
admission of live air 
for the return stroke, 
which is begun by the 
cushion air. 

In the hollow piston 
is a rifled-nut (104), en- 
gaging a rifle-bar (105). 
As the forward end of 
the piston rod is nearly 
square in section, and 
therefore cannot rotate, 




368 



COMPRESSED AIR PLANT 




'■^tzC^ 



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1- 3^ win 



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COAL-CUTTING MACHINERY 



369 



the rifle-bar is rotated at each 
stroke; and the small gear on 
the back end of the rifle-bar 
rotates the reverse valve (io6, 
109). This valve in turn oper- 
ates the spool- valve (123), by 
the dotted ports in the rear 
end of the cylinder. By a regu- 
lating valve (114), the operator 
adjusts the speed of stroke as 
the working conditions require. 
Ingersoll-Rand Pick (Fig. 
202) . The valve motion is shown 
diagrammatically by Fig. 203. 

Main ports S, S^ are con- 
trolled by slide-valve G, on the 
back of which is a lug H, en- 
gaging the spool-valve F. Air 
enters alternately the opposite 
ends of F through ports /, J^, 
which are controlled by the 
auxiliary sHde- valve K and its 
spool- valve F^. Valve F^ is 
operated by the ports A^, N^, 
N^ and N^, which connect on 
either side with main ports S, 
S^. Thus the rear end of the 
chest of F^ is connected with 
forward main port S and the 
forward end with the rear main 
port S^ ; hence, when air is ad- 
mitted to the cyhnder O through 
port S^, a smaH portion of it 
passes through N, N^ and 
throws valves F^ and K. This 
admits air from F through port 
J^ to spool- valve F and reverses 



M 



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370 



COMPRESSED AIR PLAXT 



valve G, thus opening S to live air and S^ to the exhaust. Some 
air passes from 5 through ports iV^, N^ to valves F^ and K, 
by which first K and then G are thrown forward, reversing the 
main ports and completing the cycle. The speed of F^ and K, 
and hence of F and G, is regulated by the screws L, U. 

When the piston passes the double port P the exhaust ceases. 
Therefore, if the pick misses the coal, the piston on advancing 
beyond P cushions on air in the forward end of the cyhnder. 

Stop for governor valve 

\D V Governor valve 
r 

Air Inlet 




□ 



pngO 



-Aux. valve 



T<- Regulator screV 




Fig. 203. — IngersoU-Rand Coal Pick. Diagram of Valves and Ports. 

This high-pressure air forces back the governor valve V, wholly 
or partly cutting off the inlet air to valve K, so that the machine 
runs at reduced speed until the bit again strikes the coal before 
port P is covered. The regular exhaust then takes place, the 
governor valve V opens, and the machine resumes regular 
operation. The throw of V is adjusted as desired by the stop 
D. The back stroke is begun by the cushion air confined by 
the check-valve T. Xo Hve air can pass port 5 until the piston 
has advanced far enough to reduce the cylinder pressure below 
that of the live air in 5, plus the resistance of the check-valve 
spring. The back stroke is cushioned by ports S^ and valve 5-. 



COAL-CUTTING MACHINERY 



371 



This pick is made in 3 sizes: 4I, 5^ and 6 ins. diameter. 
Maximum depth of cut, 4-6 ft.; gross weight of machine, 550- 
950 lbs. Standard wheels are 14 and 17 ins. diameter; larger 
wheels are used for shearing. 

The Hardy " coal puncher," made by the Hardy Patent 
Pick Co., Shefheld, England, is a small, hght machine, designed 
especially for driving headings, though it may be used also for 
room and longwall work. For swinging the machine on its 
mounting (usually a column), it has a worm gear, similar to 
that of the "Radialaxe" coal cutter (Fig. 204). 




Fig. 204. — Ingersoll-Rand " Radialaxe " Pick. 



Ingersoll-Rand '' Radialaxe " Pick (Fig. 204) is intended 
for shearing in driving entries, and for mining in steeply pitching 
seams where chain machines or ordinary coal picks are not 
applicable. It is an adaptation of the Temple-Ingersoll Air- 
Electric drill (Chap. XX), mounted on a column, and provided 
with a worm and wormwheel sector. A handwheel on the 
worm spindle enables the operator to swing the entire machine 
while at work, in either a horizontal or vertical arc. 



372 COMPRESSED AIR PLANT 

The group of bits, as shown, is set in a rosette socket, held 
by friction on the tapering end of the drill shank, which is 
similarly inserted in the deep socket of a long chuck. The 
individual bits are thus readily removed for sharpening and 
replacement. 

Sullivan '* Post Puncher '* (Fig. 205) is a modified rock-drill, 
designed for the same service as the Ingersoll-Rand '' Radialaxe." 
In ordinary cuts to depths of 6-8 ft., when the column cannot 
be set up close to the face, the reach of the machine is increased 
by using an extension shank. 



Fig. 205. — Sullivan " Post Puncher." At the left, the machine is arranged for 
breastwork or undercutting; at the right, for shearing. 

The bit is either solid, or of the rosette type, comprising a 
group of 5 or 7 independent, removable bits. By a segment 
and worm on the arm of the column mounting the machine is 
swung as the work requires. The cyhnder is 3 J ins. diameter; 
weight of machine, 226 lbs. 

Pneumelectric Coal Puncher, as its name implies, uses both 
compressed air and electricity. A small electric motor drives a 
pair of independent pistons in a cylinder (Figs. 206, 207). The 
rotary motion of the motor is changed to the rectilinear motion 
of the piston by the following device: The driving pinion B 
engages the horizontal gear C, which has a solid web, carrying 
a stud D (Figs. 206, 208). On Z) is a gear E, and a crank with 
crank-pin G. Within the main gear C, and attached rigidly 
to it, is a gear F, with internal teeth engaging the crank-pinion E; 



COAL-CUTTING MACHINERY 



373 




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374 



COMPRESSED AIR PL.\NT 



F having 66 and E ^^ teeth. These gears are so proportioned 
that stud D revolves in a circle concentric wdth F and of one-half 
its pitch diameter. Gear £, revoMng freely on D, causes 
crankpin G and crosshead H to reciprocate between the guides. 
To the crosshead is attached the piston rod 7, with its piston /. 
The cyHnder V contains two pistons, entirely unconnected 
^^'ith each other: the rear or dri\ing piston /, and the forward 
piston K, with its rod L and chuck X. In starting the machine 
on its first forward stroke, piston / simply pushes K forward, 
air meantime entering the cylinder behind the piston, through 




Fig. 2o8. — Pneuraelectric Coal Puncher. Diagram of Gearing. 



the valve O. On the back stroke the air m the rear of the 
cylinder is compressed and the air between the pistons is rarefied, 
thus causing K also to make its return stroke by suction, while 
air enters freely through a port at i?. At the end of the back 
stroke, the air passages below valve He between the pistons, 
whereby the charge of compressed air enters the cylinder and 
drives piston K forward on its first regular stroke. The stroke is 
cushioned after K passes the port at R. Piston K, ha\'ing 
completed its forward stroke, is followed by piston /, the air 
between them being exhausted through valve S. The return 
stroke is then made by both pistons, as at first. 

The diameter of the cylinder is 6^ ins., and the rear end 
clearance spaces are proportioned to produce a working pressure 
of 95-ioQ lbs, The motor is designed to run at three speeds, 



COAL CUTTING MACHINERY 



375 



under the operator's control, giving to the pick 140, 160 or 180 
strokes per min. It is stated that 7 H.P. are required to run 
the machine. 

The Stanley Header, originally brought out in England, is 
intended for development work in collieries, driving circular head- 
ings, for entries, airways, etc. It is now rarely used. In one of 
its forms (Fig. 209) a crosshead a, mounted on a screw shaft bj 
and carrying two horizontal arms c, cuts an annular groove, 
3-4 ins. wide. A central core is thus left, which either breaks 
up and is shoveled back as the work advances, or is blasted or 




Fig. 209. — Stanley Heading Machine for Collieries. 



wedged down from time to time, if necessary. The screw shaft 
is driven through gearing by a pair of compressed-air cylinders. 
Differential gearing produces the feed. The whole is carried 
in a frame on wheels, held firmly when in operation by jack- 
screws set against the roof. The machine is narrow enough 
to permit a man to pass alongside to the front, to throw back 
the broken coal and keep the cutter head free. 

Modifications have been introduced in this country and 
abroad. In one of them, the entire section of the heading is 
taken out in a single operation. For this, the crosshead and 
arms are replaced by a fiat, cone-shaped head, carrying a num- 
ber of individual bits, arranged in diametral lines on the cone. 



376 COMPRESSED AIR PLANT 

The circular paths traversed by the bits cover one another, 
so that the whole mass of coal is broken up. This modification 
has been used in western Pennsylvania, in some of the Frick 
Coal and Coke Co.'s mines. Average speed of advance, under 
favorable conditions, 2J-3 ft. per hour, including moving and 
setting up. In one case 2,254 linear ft. of entry were driven 
at an average speed of 17 ft. per 9 hours, and an average working 
cost of 40 cents per ft. . The exhaust assists in ventilating long 
headings. 

For rapid development work in longwall mining, the Stan- 
ley Header has been used by the Colorado Fuel Co., with the 
following results: 

Hand Labor 

2 men, i lo-hr. shift $4. 00 

Paid to men for coal produced in driving, 4I tons at 50 c. . . 2. 25 



Cost $6.25 

Distance driven in 10 hours, 3 ft. 



]\L\CHiNF. Work 

I operator, $3.00; i helper, $2.50 $ 5 . 50 

3 shovellers to load coal behind machine at $2.00 6.00 

Compressed air, repairs, depreciation and interest 3 50 

Squaring up corners, for timbering and track 5 . 00 

Cost $20 . 00 

Distance dri\'en in 10 hours, 20 ft. 



Crediting to the machine work the coal produced, viz., 15I 
tons at 50 cents loaded, the net cost for 20 ft. of entry was $1.84 
per yard. 

Auger Drills, for boring holes in coal, rock-salt and other 
soft material, are operated by compressed air or electricity. 
The Ingersoll-Rand Co. makes a breast drill, resembling a 
machinist's breast drill. It has a 3-cyhnder motor, which can 
readily be reversed for withdrawing the bit from the hole; weight, 
exclusive of the bit, 18 lbs. Another heavier machine is mounted 



COAL-CUTTING MACHINERY 377 

on a column or bar. A compressed air auger mounted on single 
or double column, is made by the Jeffrey Manufacturing Co.; 
weight, for a 6-ft. vein, 183 lbs. Speed of the engine is about 
3,000 revs., and that of the feed shaft, 850 revs, per min. 

Auger drills operated by compressed air, electricity or hand 
power, are also made by the Ho wells Mining Drill Co., Ply- 
mouth, Pa. 

The Fairmont Mining Machinery Co., Fairmont, West Va., 
makes an auger drill for mounting on electric-driven chain coal 
cutters. 

Comparison of Coal Cutters. The chain machines, which 
have the widest application, work best in clear coal, of uniform 
quaHty, though the recent types are suitable also for hard, 
" bony " coal, or coal containing streaks of pyrites ('' sulphur 
balls"). In coal of this character, the pick machines work well, 
because the operator can regulate the strength of the blow and 
so direct the machine as to cut around a hard place. 

Somewhat less slack and fines are made by the chain machines, 
as the volume of undercut is smaller; but, when the product 
goes to coke ovens, the larger quantity of fines made by pick 
machines is immaterial. Also, in solid, hard coal, the higher 
undercut of the pick machines causes a more complete breaking 
up of the whole mass when blasted down, and the coal, therefore, 
is sometimes more readily loaded. 

For chain cutters a fairly good roof is desirable; otherwise 
props may have to be set so close to the face as to interfere 
with the manipulation and shifting of the machines. Chain 
machines can be worked in seams as thin as about 20 ins., 
though they are more conveniently operated in thicker seams. 
The continuous-feed chain and disk machines are specially 
useful for thin, pitching seams in longwall work, as they operate 
with almost equal facility either up or down the pitch. Disk 
cutters are now rarely used. 

The " mining rate," or cost of mining by hand, together with 
the character of the coal seam, will usually determine whether 
coal cutters can be economically applied in a given mine or 
district. In general, for seams of average quality and thickness, 



378 



COMPRESSED AIR PEAXT 




CJ 



o 



< 



COAL-CUTTING MACHINERY 379 

when the local cost of hand mining is not less than 55-60 cents 
per ton, a saving may be effected by introducing machines.* 

Loading Machines of several makes are used underground 
to some extent for loading cars in the working places of both 
coal and metal mines. Fig. 210 shows a compressed-air driven 
loader of the conveyer type, for collieries, which may be men- 
tioned here in connection with coal-cutting machinery. A con- 
siderable number of them are now employed in the United States 
and Canada. The frame of the motor supports the elevated 
end of the conveyer, near the car. 

* This applies to pre-War conditions. 



CHAPTER XXIII 
CHANNELING IMACHINES 

Originally, channeling machines were used almost exclu- 
sively for getting out dimension stones in quarry work. Of late 
years, however, they have been employed in increasing numbers 
for certain kinds of rock excavation, where it is desired to have 
smooth, uniform walls; for example, rock cuttings for railroads, 
canals and water-wheel pits for power plants. They are best 
adapted for cutting the softer rocks, like limestone, most of the 
sandstones, slate, shale, etc., though they may be used also for 
some of the varieties of granite, gneiss, porphyry, schist, and 
other metamorphic rocks. Hard rocks are best quarried by 
drilling rows of holes, with wedging or blasting. 

In a certain sense channelers resemble reciprocating rock- 
drills, a single bit or a " gang " of bits being attached to the 
piston rod. But, instead of drilhng a series of round holes, the 
channeler, as its name implies, cuts a continuous, narrow groove, 
without rotation of the piston and bit. In its typical form, 
the machine is solidly supported on a heavy carriage or truck, 
generally mounted on a track laid along the Hne to be cut. The 
motive power may be compressed air, electricity, or steam. By 
means of an auxihary engine and worm gearing, the whole 
machine, while at work, is fed forward automatically at a suitable 
speed. Fig. 211 shows the general construction of a standard 
compressed air-driven channeler. Another design, a track chan- 
neler for cutting marble, with adjustable mounting and a re- 
heater mounted on the carriage, is shown in Fig. 212. 

General Construction. The construction of channeling 
machines is varied to suit the conditions of work: i. For cutting 
vertical channels only, the rigid head machine is used; that is, the 
standard supporting the cyhnder and accessories is non-adjust- 

380 



CHANNELING MACHINES 



381 




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c 

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382 



COMPRESSED .\IR PLANT 




C 

c 

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CHANNELING MACHINES 



383 



able, being permanently fixed in an upright position. Fig. 213 
shows a steam-driven channelcr of this class. It is employed for 
canal and railroad cuttings, general rock excavation and for 




Fig. 213. — Sullivan Rigid Back, Steam-Driven Channeler. 



quarrying where the strata are horizontal or nearly so. 2. For 
quarrying building stone lying in inclined stratified beds, like 
most limestones, the channels must generally be cut at right 
angles to the bedding planes; hence, the mounting of the cutting 



384 COMPRESSED AIR PLANT 

engine is adjustable, for making a channel at any desired angle 
to the vertical. The cylinder with its appurtenances is swivelled 
on its supporting frame or standard; or the frame may be pro- 
vided with T-slots (resembling those of the table of a planer), 
by means of which the cyhnder is bolted firmly in the required 
position. The supporting frame, in turn (see Fig. 212, of the 
Ingersoll-Rand Ram Track Channeler), may be swung back to 
a nearly horizontal position, for making wall cuts along a quarry 
face. In different designs, the minimum swing-back angle varies 
from 15° or 20° to i,t,° to the horizontal ; ^ but it is not often 
necessary to cut with these machines at less than 45°. Fig. 214 
shows a Sullivan channeler of this class. 3. A third form, 
designed specially for making horizontal channels, or '* under- 
cuts," is used less frequently than the others. An Ingersoll- 
Rand machine of this type is shown in Fig. 215. As indicated 
in the cut, the head may be bolted to either end of the carriage. 
4. Lastly, a light machine, which is in effect a large rock-drill, 
may be mounted on a '' quarry bar " — a long, hollow bar, 
supported at each end by a pair of inclined legs.* This is 
generally used for drilling a row of holes placed close together, 
the partitions between which are afterward cut out by a "broach- 
ing " bit. A modification of the quarry-bar machine is made 
by the Ingersoll-Rand Co. It is a true channeler, mounted on 
a heavy swivel plate, which slides on a pair of horizontal bars, 
about 10 ft. long, supported by inclined legs (Fig. 216). The 
whole machine is fed along the bars automatically by a small, 
3-cylinder engine, which actuates a travelling feed-nut, engaging 
with a threaded shaft between the bars. 

There are many variations in construction of the above- 
mentioned classes of channeler, to adapt them to local condi- 
tions. Among other machines, of entirely different design, 
may be mentioned the Wardwell and the Bryant channelers. 
The Wardwell, a heavy machine operated by steam only, 
has been in successful use for many years. It is intended 
for making vertical channels, a gang of bits being set in a 

* In one of the Sullivan models, the legs are replaced by vertical standards, 
each carried on a small wheeled truck. 



CHANNELING MACHINES 



385 



massive frame, which is given an up and down movement, 
something like a jumper drill. 

For the first three classes of channeler a gang of from 3-5 
bits is employed. These have long square shanks and are set 
closely side by side, the cutting edges being alternately at right 




Fig. 214.— Sullivan Adjustable Back Air-Driven Channeler. 



angles and at 45° to the direction of the channel. This arrange- 
ment forms practically a succession of Z-shaped bits, and insures 
the cutting of a regular channel with smooth walls. The bits 
are clamped firmly in a heavy chuck, attached to the piston rod 
of the engine, and are guided either by a crosshead or (as in some 



386 



COMPRESSED AIR PLANT 




CHANNELING MACHINES 



387 




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388 COMPRESSED .\IR PLAXT 

of the IngersoU-Rand patterns) by a pair of roller guides. Some 
sa\ing in power has been realized by the introduction of the roller 
guides; they ehminate part of the weight due to the crosshead, 
which must be hfted at each stroke, and the friction loss is 
reduced. The SuUivan Company builds a duplex or double- 
head machine. There are two cylinders, side by side on a heavy 
frame, each with its gang of bits and operated by a single valve- 
chest. As the blows alternate, one piston making its down 
stroke while the other is on the up stroke, the machine can be 
run at high speed without excessive \'ibration. Its working 
capacity is correspondingly greater than that of the single- 
cylinder machines. When the plant consists of a few machines 
only, they may be advantageously driven by steam (Fig. 213); 
but. for large-scale work, a higher degree of economy results 
from the emplo}Tnent of compressed air. furnished by a central 
plant. Each machine is then pro^'ided with its o\sti reheater,* 
mounted on the carriage Fig. 212). Air pressures generally 
range from S5-110 lbs, 

^^Tlile at work the main cyhnder of the channeler is raised 
and lowered in its guide shell by a screw-feed, operated auto- 
matically or by hand. The hand feed is rarely employed except 
for the smaller machines. The automatic feed may be caused 
either by an independent engine, similar to that used for the 
longitudinal feed of the quarr}--bar machine, already referred to; 
or, b}- a chain and sprocket drive from the machine which 
furnishes the propelling power along the track. The chain feed, 
as used in the Sulhvan channelers, is shown in Figs. 211 and 213. 
Most of the Ingersoil-Rand channelers are provided vdih the 
independent feed engine, which is of the 3-cylinder 1)^6, ver>^ 
small and compact in design. In either case, when the cut has 
reached the required depth, the feed is reversed and the entire 
head, with its accompanying parts, is raised preparatory to 
making the next cut. 

Depth of Cut and Speed of Work. The hea\'iest channelers — 
those with rigid back or standard — will cut to depths of from 8 

* See Chap. XIX, on Reheaters. 



CHANNELING MACHINES 389 

-15 or 16 ft., according to the character of the stone; the swing- 
back and bar machines will cut from say 6-10 or 12 ft., and 
undercutting machines up to 7 ft. For starting a channel, the 
width of a bit is from ij to a maximum of 4 ins., depending 
on the depth of cut to be made and on the nature of the 
stone. The gages of the successive bits are generally rfeduced 
by j^ in. each, the finishing bits usually cutting a width 
of i| in. 

' The cutting capacity of channelers varies greatly. It is 
largest in the softer stones, when of uniform texture and quality, 
and in fully developed quarries, where the work is systematic 
and the stone hes below the zone of weathering and surface 
disintegration. In sandstone of average hardness and under 
favorable conditions, from 250-300 sq. ft. of channel may be cut 
per 10 hours by the heavy machines; or, including all stoppages 
and delays, from 4,000-4,500 sq. ft. per month; in the softer 
sandstones and limestones higher duties are obtainable. The 
swivel-head and other adjustable channelers are lighter than 
the fixed-back machines and in the same kind of stones their 
rate of work is generally slower. Machines working in rather 
hard marbles, hke those of Rutland, Vt., will cut from 2,300- 
2,500 sq. ft. per month, or an average of 85-100 sq. ft. per day. 
A single day's work, however, will often greatly exceed these 
figures. In hard marble or limestone, the smaller bar machines 
will cut an average of say 40 sq. ft. per 10 hours and up 
to 125 sq. ft. in softer stones. For hard gneiss, or schist, 
like that of New York island, an average duty would be 
65-70 ft. per day. 

Tables XL and XLI, showing dimensions, weights, and other 
data, of the channelers of two well-known builders, will further 
illustrate the features of these machines. 

The Ingersoll-Rand Company have applied the principle of 
their '' Air-Electric " rock-drill to the design of the cylinder 
and air compressing mechanism of a track channeler (Fig. 217). 
That is, an electric motor, mounted on a carriage, drives a sin- 
gle-acting air compressing pulsator, which is connected to the 
channeler cylinder in a manner similar to that of the Temple- 



390 



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392 



COMPRESSED AIR PLANT 




CHANNELING MACHINES 393 

Ingersoll rock-drill, described in Chap. XX. For each double 
stroke of the pulsator, there is a blow and return stroke of the 
channeler piston, the speed of stroke being thus controlled 
by the speed of the electric motor which furnishes the power. 
Favorable results, both as to power cost and maintenance, 
have been secured. This channeler has a swivel head and a 
swing-back support, and is therefore suitable for varied quarry 
service. 



CHAPTER XXIV 
OPERATION OF MINE PUMPS BY COMPRESSED AIR 

It is intended here to deal only with that part of the extensive 
subject of mine drainage which has to do with the employment 
of compressed air as a motive power. Under this head there are 
three general forms of apparatus: 

1. Direct-acting pumps: single-cylinder, duplex, or com- 
pound. 

2. The air-Hft pump. 

3. Pneumatic displacement pumps. 

In this chapter the first class only will be considered. 

Simple, Direct-Acting Pumps. Notwithstanding the general 
similarity in the behavior of steam and compressed air, when used 
in the cylinders of direct-acting pumps, there are some important 
points of difference. By first considering briefly the construction 
of the types of pump in common ase the results obtainable from 
the employment of compressed air can best be set forth. 

The development of the direct-acting pump dates from 
Henry R. Worthington's invention in 1841; and a large part 
of all the pumping in the mines of this country, and much of it 
in other countries also, is done by pumps of this class. The 
cylinders are set tandem, the power being transmitted from the 
steam to the water cylinder through a piston-rod common to 
both. As there are no rotating parts, the length of stroke is 
controlled by the admission and exhaust of the steam. In all 
the simple pumps the valve motion involves the use of an 
auxihary valve, the movements of which are governed by the re- 
ciprocating movement of the piston, and which in turn operates 
the main valve. The duplex form consists essentially of two 
simple pumps, set side by side, with an interdependent valve 
motion; that is, the valve of each is operated positively, through 

394 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 395 

a system of levers, by the movements of the piston of the other 
side. 

Though direct-acting pumps are strong and rehable, simple 
in construction, and occupy but little space, they are extremely 
uneconomical machines, unless the steam cylinders are com- 
pounded. It is hardly necessary to say that this ought not to be 
the case. Pumping is an operation that should be conducted 
economically, especially in connection with mining, where the 
pumping of water is classed as " dead work." Moreover, the 
conditions in themselves are not unfavorable. A pump works 
under a practically constant load, from the beginning to the end 
of each stroke, the only necessary variation — which need 
not be large — occurring at the instant the discharge valves 
open. 

The trouble is that, in attaining compactness, simplicity, and 
moderate first cost, the power is not applied in simple, direct- 
acting pumps to the best advantage. As there is a constant 
load, but no fly-wheel to equalize the power, steam must be 
admitted at full pressure throughout the entire stroke ; otherwise 
the piston would be unable to reverse, and would come to a 
standstill. Such a pump must work practically without cutoff, 
and therefore a cylinderful of steam, nearly at initial pressure, is 
exhausted at each stroke. In some pumps the terminal pressure 
is quite as high as the initial. A duplex, non-compound pump, 
having a positive valve motion, may at times be even a more 
extravagant steam-consumer than a single-cylinder pump, since 
one piston may reach the end of its stroke before the other is 
ready to reverse its valve. In such case the momentum of the 
incoming steam fills the cylinder at initial pressure at the moment 
of exhaust. 

For steam-driven pumps there are several ways of improving 
these conditions: 

I. The adoption of compound or triple expansion cylinders. 
This type is suitable for the larger sizes of pump, and its use is 
increasing for mines where the depth and quantity of water 
warrant the higher first cost. The space occupied is little greater 
than for simple pumps of the same capacity, and satisfactory re- 



396 COMPRESSED AIR PLANT 

suits are obtained when they work under proper conditions and 
with sufficient initial pressure. 

2. While retaining the tandem form, a fly-wheel may be intro- 
duced, driven from the crosshead or from the steam-cylinder 
connecting-rod. This is a reversion to a type of pump long ago 
discarded for general service in this country, in favor of the 
simpler but less efficient form Vv^ith no rotating parts. Although 
such a pump occupies much more space and its first cost is 
increased, there can be no doubt as to the advantages of being 
able to use the steam expansively, without the necessity of com- 
pounding. A large number of pumps of this description are now 
employed in mines ; many of the Riedler pattern and some of less 
elaborate and expensive design, such as the Prescott and others, 
in which an early cutoff — at one-quarter or even one-eighth 
stroke — is satisfactorily adopted. 

Notwithstanding the advances made along these lines in the 
mechanical engineering of pumps and the added economy gained 
in their operation, it has been very generally assumed in the past 
that similar economies are not attainable when compressed air 
instead of steam is employed as the motive power. Yet the 
advantages accruing from the utilization of compressed air trans- 
mission in mines are marked. As the heavy losses due to radia- 
tion and the condensation of steam in pipe-lines are avoided, the 
transmission of power by compressed air may be conducted with 
a high degree of efficiency. No difficulty exists as in the disposal 
of exhaust steam underground, nor is any danger to be appre- 
hended from the rupture of a compressed-air pipe, while the 
bursting of a steam pipe in a shaft or in the mine workings may 
cause serious trouble. The failure to realize these advantages, 
and the unsatisfactory results obtained in most cases from com- 
pressed-air-driven pumps, are due largely to the fundamental 
differences in the behavior of steam and compressed air when 
used in a motor cylinder. In Chap. XVII reference has been 
made to the reduction of cylinder temperature accompanying 
the expansion of compressed air. The point of cutoff being the 
same, this causes lower terminal and mean pressures with air 
than with steam. In other words, at a given initial pressure 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 397 

and without reheating, a cyhnderful of air develops less 
power. 

This property of air, together with the fact that it does not 
condense, indicates clearly that steam and compressed air are not 
equally well adapted for use in an engine of the same design. It 
is not easy to understand, therefore, why mechanical engineers 
and especially pump-builders have not given more attention to 
the production of pumps properly designed for the use of com- 
pressed air. Few, if any, other branches of motor-engine prac- 
tice have been so neglected. Lack of information among users 
of compressed air is responsible in part ; in addition to which it is 
not generally realized that relatively unimportant modifications, 
at small cost, would produce much better results. Users of the 
ordinary steam pump have become accustomed to its low econ- 
omy, and, because it is strong and serviceable, it is apt to be 
accepted without question when compressed air is used instead of 
steam. But in applying compressed air to the inefficient single- 
cylinder pump, as usually designed for steam, the net result is no 
better and may be even worse, than that obtained from steam. 
The clearance spaces are large and, as the air is admitted to the 
cylinder throughout full stroke, it is used in a wasteful manner. 
Moreover, the stroke is often shortened by imperfections in the 
valve action. 

Another unfavorable feature of mine pumps driven by com- 
pressed air is the frequently improper selection of the cylinder 
proportions and arrangement of the plant. In mines having a 
number of levels the pumps are distributed according to varying 
requirements as to height of lift and quantity of water to be 
raised. The lowermost pump may have to work under a heavy 
head; others under a head of only loo or 200 ft. As all are 
usually operated from the same pipe-line and under a common 
air pressure, it is clear that the dissimilarity of working conditions 
must be met by proportioning the water and power ends of each 
pump according to the work to be done. But, through error 
or carelessness, the power end is often badly out of proportion, 
the tendency being to err on the side of furnishing too much 
power. The steam (or air) cylinder may be of such size as to 



398 COMPRESSED AIR PLANT 

require a pressure of only 30 or 40 lbs. per sq. in., while the pipe- 
line pressure is 70 or 80 lbs., as usual with mine compressor 
plants. So it often happens that the deepest pump in the mine 
is the only one operating under a proper pressure. The cylinders 
of the others, even if running under throttle, are filled wdth air 
at full pressure when exhaust takes place.* 

The difficulty with common direct-acting pumps is thus two- 
fold : the air is used without expansion, and the pressure is often 
higher than is necessary. Recognizing, however, the convenience 
with which the inexpensive, ready-made single-cylinder pumps 
may be installed, and that in many cases efficiency of operation is 
really a secondary consideration, a few points will here be dis- 
cussed as to their employment, and the volume of air required 
for a given quantity of work. Questions relating to the expan- 
sive use of compressed air for pumps will be taken up afterward. 

Cylinder Dimensions of Simple Pumps. In calculating the 
sizes of cylinders for a simple, or single-cylinder pump, to work 
under given conditions, the dimensions of the water cylinder 
must first be determined. There are three variables to be 
dealt with, viz., diameter, length of stroke, and number of strokes 
per minute; or the last two factors named may be combined in 
the shape of piston speed per minute. The volume of water to 
be raised being given, the cy Under dimensions may be obtained 
from lists of standard sizes of pumps, which would usually 
be adhered to on the ground of saving in first cost. With a given 
air pressure and head of water, the diameter of the air cylinder 
obviously depends upon that of the water cylinder. The follow- 
ing relation between the two has been determined by Mr. William 
Cox:t " Area of air cylinder is to area of water cylinder as half 
the head is to the air pressure." 

In using Table XLII, ratios for intermediate heads and pres- 
sures may be obtained by interpolation. 

In this table the unit diameter of water cylinder is taken as 
I in. Diameters of air cylinders, as calculated, will be in deci- 

* Some suggestive remarks on this subject are made by Frank Richards, 
" Compressed Air," pp. 171-172. 

t Compressed Air Magazine, Feb., 1899, p. 583. {^y permission.) 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 



399 



mals, and often of odd sizes not occurring in practice. After 
determining the exact diameter, the nearest standard diameter of 
cyHnder would be chosen and the air pressure and piston speed 
adjusted accordingly. 

Table XLII 

Ratios or Diameter of Air Cylinder to Diameter of 
Water Cylinder. (Williajvi Cox) 





Air Pressure, Lbs. 


Head in Ft. 


















20 


25 


30 


35 


40 


45 


50 


50 


1 . 12 


1 .00 


0.91 


0.84 


0.79 


0.74 


0. 71 


100 


1.58 


1.41 


I. 29 


1 . 20 


1 . 12 


105 


1. 00 


125 


1.77 


1.58 


1-45 


1-34 


1-25 


1. 18 


1. 12 


ISO 


1.94 


1-73 


1-58 


1-45 


1-37 


I. 29 


1 . 22 


175 


2.09 


1.87 


1 . 70 


1.58 


1.48 


1-39 


1.32 


200 


2. 24 


2.00 


1.82 


1 .69 


1.58 


1.49 


1. 41 


225 


2.37 


2. 12 


1.94 


1.79 


1.68 


158 


I 50 


250 


2.50 


2. 24 


2.05 


1 .90 


1.77 


1 .67 


1.58 


27s 


2.62 


2-35 


2.14 


1.98 


1-85 


1-75 


1.66 


300 


2.74 


2.45 


2.24 


2.07 


1.94 


1.82 


I 73 


325 . 


2.8s 


2.55 


2- a 


2. 16 


2.02 


1 .90 


1.80 


350 


2.96 


2.64 


2.42 


2.24 


2.09 


1.97 


1.87 


375 


3.06 


2.74 


2.50 


2.31 


2. 16 


2.04 


1.94 


400 


3.16 


2.83 


2.58 


2-39 


2.23 


2. II 


2.00 


42s 


3.26 


2.92 


2.66 


2.46 


2.30 


2.17 


2.06 


450 


3-35 


3.00 


2.74 


2-53 


2-37 


2.24 


2. 12 


475 


3-44 


3.08 


2.82 


2.60 


2.44 


2.30 


2.18 


500 


3-53 


3.16 


2.89 


2.67 


2.50 


2.36 


2. 24 



Volume of Air for Pumps Working without Expansion. To 

determine the volume of free air for a single-cylinder pump, 
use the following formula:* 

V = 0.093 W2 — ^— , in which: 

V = volume of air in cubic feet per minute; 

h = head in feet under which the pump is to work ; 

* Compressed Air Magazine, Feb., 1899, p. 581. (By permission.) 



400 



co:mpressed air pl.\nt 



G = gallons of water to be raised per minute ; 
P = receiver gage pressure of air to be used ; 
\V2 = volume of free air corresponding to i cu.ft. at the given 

pressure, P. 
In this formula, i59c has been added to the volume of air 
to cover losses. The following table gives values of W2 and 
0.093 W2 for different pressures: 

Table XLIII 



Air Pressure. P. in Lbs. 


W: 


0.093 W: 


15 


2.02 


O.1S786 


20 


2.36 


0. 21948 


25 


2.70 


0. 251IO 


30 


3 04 


0. 28272 


35 


338 


0.31434 


40 


3-72 


0.34596 


45 


4.06 


0.37758 


50 


4.40 


0.40920 


55 


4-74 


0.44082 


60 


5.08 


0.47244 


65 


542 


0.50406 


70 


5 76 


0.53568 


75 


6. 10 


0.56730 


80 


6.44 


0-59890 


85 


6.78 


0.63054 


90 


7 I - 


0.66216 



For example, let it be required to find the volume of free air 
per minute required to raise 200 gals, of water to a height of 
150 ft., the gage pressure being 30 lbs. From the table, 0.093 
W2, corresponding to 30 lbs. =0.2827; hence, 

200X1 so 

V = 0.2827 X ^— = 282.7 cu.ft. free air. 

30 

The horse-power may be calculated from Table XLIV, in 
which the mean pressures per stroke (from Table \TI, Chap. X). 
for the different terminal pressures, are given in the second 
column, and the horse-powers in the third column: 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 401 



Table XLIV 



Terminal Pressure, Lbs. 


Mean Pressure per Stroke. 


H.P. per Cu.ft. Free Air. 




20 


14.40 


0.0628 




25 


17 


01 


0.0743 




30 


19 


40 


. 0847 




35 


21 


60 


0.0943 




40 


23 


66 


0.1033 




45 


25 


59 


0. III7 




50 


27 


39 


O.II96 




55 • 


29 


II 


0. 1270 




60 


30 


75 


0.1340 




65 


32 


32 


0. 1406 




70 


33 


83 


0. 1468 




75 


35 


27 


0.1527 




80 


36 


64 


0.1583 



As the horse-power corresponding to a given terminal pres- 
sure does not increase in constant ratio with the initial air 
pressure, it follows that the higher pressures are not so economical 
for simple pumps as low pressures. Expressed in another way, 
the work of compression decreases with the air pressure, and 
therefore the useful work done in a pump using air at full pressure 
is greater at low pressures and its efhciency is increased. Thus, 
in the example given above, the horse-power developed in using 
the 282.7 cu.ft. of free air, at a pressure of 30 lbs., is: 

282.7 X0.0847 = 23.94 H.P. 

If the air pressure employed were 50 lbs., the cu.ft. of free 
air would be 245.52 and the corresponding H.P., 29.36, the 
added power cost being 5.42 H.P. It may be stated that the 
difference in favor of the lower air pressure is offset in part by the 
fact that, at the higher pressure, a pump with a smaller power 
cylinder will do the same work, thus saving in the first cost. 

But the low pressures thus shown to be suitable for simple 
pumps would not serve for machine drills, which must be con- 
sidered first, as they are in nearly all cases the chief users of com- 
pressed air in mines and quarries. To secure the best results 
from the pumps, a separate, low-pressure compressor would be 



402 



COMPRESSED AIR PLANT 



required, a provision which is usually out of the question. Since 
it is generally necessary to use high-pressure air, at, say, 80 or 
90 lbs. gage, the air must either be wire-drawn into the pump 
cylinder or else reduced to the required pressure before being 
delivered to the pump. 

In the first case, the results as to volumes of air used, as given 
in the preceding discussion and tables, must be modified by 
introducing a factor of increase, based on the ratio which the 
pressure to be used in the pump bears to the pressure carried 
in the air main. Edward A. Rix furnishes a table,* part of which 
is abstracted in Table XLV. It shows the volumes of free air 
theoretically required for a unit of 10,000 ft.-gals. of work 
( = 83,000 ft. -lbs. or 2.5 H.P.), at different air pressures, referred 
to a standard receiver pressure of 90 lbs. 

Table XLV 



Gage 

Pressure. 

Lbs. 


Ratio of 

Compres- 
sion, Re- 
ferred to 
90 Lbs. 


Cu.ft. of Air 

Calculated 

from Cox's 

Formula. 


Factor of 
Increase for 
Wire- Draw- 
ing from 
90 Lbs. 


Increased 

Volume, 

Cu.ft. 


Actual H.P. 
at 90 Lbs. 


Efficiency 
on Basis of 

2.5 H.P. 
Theoretical. 


20 


3- 


113 


1.26 


142 


28.6 


9 


25 


2 


6 


108 


I . 22 


125 


25 




10 


30 


2 


3 


97 


I. 19 


115 


23 




II 


35 


2 


I 


93 


I. 17 


108 


21 


5 


II. 6 


40 




9 


89 


I. 14 


102 


20 


5 


12. 2 


45 




7 


87 


I . 12 


97 


19 


7 


12.7 


50 




6 


85 


I . II 


93 


19 




13 I 


55 




5 


82 


I .09 


89 


18 


2 


137 


60 




4 


80 


1.07 


86 


17 


4 


14 3 


65 




31 


79 


I .06 


84 


16 


8 


14.9 


70 




24 


78 


I 05 


82 


16 


4 


15-3 


75 




17 


77 


1.04 


80 


16 




15-6 


«o 




I 


76 


I 03 


78 


15 


6 


16. 


85 




05 


75 


I .02 


76 


15 


2 


16.4 


90 







74 


I .0 


74 


14 


8 


16.9 



The factors in column 4 are assumed as about 70% of the 
ratios of the absolute temperatures due to expansion of the 
air from 90 lbs., to the air pressures in column i. They may be 

* Transactions Technical Society of the Pacific Coast, Aug. 3, 1900. 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 403 

taken to apply when the length of air main from the compressor 
to the pump is moderate, as in carrying the air to a pump situated 
at the bottom of an ordinary shaft. The showing is a poor one, 
but the unfavorable working conditions, as to the type of pump 
and mode of using the air, must be taken into account. 

In the second case, the normal air pressure carried in the 
mine (say, 90 lbs.) may be reduced to a suitable pump pressure 
by placing a reducing valve in the air main. The increase 
of volume thus produced will be accompanied by a considerable 
drop in temperature, so that the full increase is not realized. 
Part of the lost heat will be regained by friction, and from 
external sources if there be any considerable length of pipe 
between the reducing valve and pump; but the efi&ciency will 
be materially increased if the cold, partly expanded air be 
passed first into an underground receiver and thence to the pump. 
This arrangement has been satisfactorily adopted, for example, 
in the case referred to at middle of p. 256. An adjustable spring- 
reducing valve is set to furnish any desired pressure below that 
in the main. That is, the volume of air allowed to pass is such as 
to maintain automatically a certain difference in pressure 
between that in the main and the pipe leading to the second 
receiver. The latter serves three purposes : (i) if it be of ample 
size or of the tubular type the air will regain nearly, if not quite, 
its normal temperature; (2) much of the entrained moisture 
will be deposited, and trouble from freezing avoided; and (3) 
the receiver, if placed near the pump, will minimize the pulsa- 
tions and equalize the air pressure. 

In the particular instance to which reference is here made, 
two underground receivers were installed, 300 ft. apart, the 
reducing valve being put in the main just above the first receiver. 
This arrangement not only caused a very complete deposition 
of the moisture, but the air entirely recovered its normal tem- 
perature by the time it left the second receiver on its way to 
the pump. The main air pressure was 85 lbs., and at the pump 
about 45 lbs. Indicator diagrams showed 128.5 H.P. developed 
by the compresssor and 16.45 H.P. at the pump, or an efhciency 
of 12.5%; thus agreeing quite closely with the figures in Table 



404 COMPRESSED AIR PLANT 

XLV. Subsequently, by compounding one of the pumps, 
using 62 lbs. initial pressure in the high-pressure cylinder and 
admitting some live air to the intermediate pipe between the 
cylinders, the efficiency was raised to 25.9%. This must be 
considered a fairly satisfactory performance for a pump not 
specially designed for its work. 

By adopting stage compression or by reheating, or both, the 
total efficiency can of course be increased considerably beyond 
the efficiencies show^n in the table. Mr. Rix states that by 
actual test of a number of simple pumps he found their work to 
be approximately 135 ft. -gals, per cu.ft. of free air. For stage 
compression the efficiency is increased by 15% (giving, say, 155 
ft. -gals.), and, by reheating, the 135 ft. -gals, is increased by 
the ratio of the absolute temperatures under which the pump 
works, without deducting the small cost of reheating. 

Prevention of Freezing of Moisture. Though this subject 
has already been discussed at some length, several additional 
points may be noted in connection with pumping. Some benefit 
may be derived by leading a jet of water from the pump column 
into the air pipe, just before reaching the pump. A very sm.all 
quantity of water will suffice to prevent an excessive drop in the 
temperature of the exhaust. A better way is to tap a i-in. pipe 
into the column pipe, draw down the end of this pipe to, say, 
one thirty-second of an inch and insert the nozzle so formed 
into the exhaust port. The author has observed the plan of 
carrying a small steam jet close to the exhaust port; but it is 
obvious that this is feasible only when steam is used near by for 
some other purpose. Moreover, steam so applied is utilized 
much less perfectly than when used in a cylinder jacket. If 
steam be available, a little may be injected into the feed air pipe 
near the pump. An intimate mixture between the steam and 
air is thus produced, and in condensing the latent heat of the 
steam is given up. If water at 212° F. be injected, each pound 
in cooling down to 32° F. will give up 180 thermal units. But 
with steam at the same initial temperature, each pound in con- 
densing gives up 966 thermal units, in addition to the 180 units 
imparted in cooling to 32°. Still another mode of preventing 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 405 

freezing is to warm the compressed air by passing it through a 
coil of pipe, placed in an enlarged section of the water column, 
or else in the pump suction pipe. 

Compressed-Air-Driven Compound Pumps. It is a commonly 
held idea that if compressed air be used for operating compound, 
direct-acting pumps, it should be employed like steam, With a 
cutoff in each cylinder. The resulting drop in cylinder tempera- 
ture would obviously be less than that caused in a single cylinder 
by the same ratio of expansion from a given initial pressure. 
But in aiming thus to attain a higher efficiency, by adopting the 
largest possible range of expansion, very low cylinder tempera- 
tures would still be produced. The loss of heat takes place 
chiefly within the cylinder, instead of in, and just outside of, the 
exhaust port, as is the case with pumps working at full pressure. 
Furthermore, though the same total fall of temperature occurs in 
either case, when the air expands within the cylinder the force 
of the exhaust is diminished by the low terminal pressure, and 
the ports are the more liable to be choked with ice. 

In order to use the air expansively the necessity for reheating 
in some form is clearly indicated, aside from any question of gain 
in economy. Various plans have been tried of warming the 
cylinders by the application of external heat, such as developing 
them in hot-air jackets, surrounding them by water, even heat- 
ing them by the flamxes of large lamps or torches. But air is 
too poor a conductor of heat to render these means efficient. 

The mode of applying extraneous heat may be varied in 
several ways, viz., (i) Preheating the compressed air sufficiently 
to permit of a reasonably early cutoff in each cylinder, while still 
avoiding too low an initial temperature in the low-pressure cylin- 
der; (2) in addition to preheating, the air may be reheated 
between the cylinders; (3) using cold air at full pressure in the 
high-pressure cylinder and expanding into the low-pressure 
cyhnder, with or without reheating; (4) using cold air at full 
pressure in both cylinders, the air being expanded between them, 
with the application of reheating. 

The first two methods are feasible when the compound pump 
is of suitable design and the heating properly apphed; but there 



406 COMPRESSED XIR PL.\XT 

would be an undesirable variation in power and speed, for an 
engine necessarily working under a constant load, if the pump be 
of the usual direct-acting t^-pe. ^\'ithout fly-wheel, ^loreover, 
under the first plan a high initial temperature would be necessary. 
If the expansion be adiabatic. from an initial pressure of, say, 
80 lbs. to atmospheric pressure and normal temperature, the 
temperature to which the air would have to be preheated is 
given by the expression: 

T' = t(^)'^ or. T' = 70°+459^(^')°" = 446= F. 

Although this temperature would be rapidly lowered during 
the stroke, proper lubrication of the cyUnder might be interfered 
^-ith. The third method would avoid in part the difficulty of 
variation in power and speed, though there would still be a vari- 
able back-pressure on the high-pressure piston; but the increase 
in volume due to clearance, and on expanding into the passages 
and intermediate pipe to the low-pressure cyhnder, would con- 
siderably reduce the temperature of the air. and a large further 
drop would ensue during the work of expansion in the low-pres- 
sure cylinder. Such temperature drop may be prevented, or at 
least diminished, by introducing a receiver-reheater between the 
cylinders, "^ith material gain in efficiency. This method has fre- 
quently been adopted, and on the whole is much preferable to the 
two first mentioned. 

The fourth arrangement, however, appears to be the most 
satisfactory. As has been pointed out by E. A. Rix,* in the 
practical appHcation of compressed air to pumps only a small 
part of the total possible work of expansion \\'ithin the two 
cylinders can be realized, even in favorable circumstances. 
Nevertheless, if properly installed and operated, it becomes 
perfectly practicable to drive a compound pump by compressed 
air. It is a much more satisfactory machine than a single- 
cyhnder pump, and is capable of working with a fair degree 
of efficiency. This may be accomplished by expanding the 

* Transactions Association of Engineering Societies, 1900. Mr. RLx also pro- 
poses the use of three-, and even four-cylinder pump)s. 



OPERATION OF MINE PUMPS BY COMPRESSED AIR 407 

air between the cylinders only, restoring the consequent loss of 
pressure by reheating and employing full pressure in both 
cylinders. Thus no drop of temperature takes place in the 
cylinders themselves, and the pressures, back-pressures, and 
speed are constant. Each air card is practically rectangular 
in shape. The pressure drop between the cylinders may be 
made small; in fact, it need not be more than is sufficient to 
give the head necessary to cause an active flow of air into the 
intermediate reheater and thence to the low-pressure cyHnder. 
A drop of, say, 20 lbs. for an initial pressure of 70-80 lbs. will 
usually answer. 

The degree of heat to be imparted by the intermediate 
reheater, to restore the heat lost by a drop of 20 lbs., would be 
only 204° F., for a final temperature of 60° at exhaust. If the 
pump be suitably situated, an ordinary fuel-burning reheater 
may be employed; or, should this be inadmissible, the water 
ftom the pump suction or column pipe may be utihzed for 
reheating, as already suggested. An example of this arrange- 
ment, which has often been cited, is to be found in the Gwin 
Mine, Calaveras Co., California. A Worthington compound 
pump, having a capacity of 200 gals, per min., was installed 
on the 600-ft. level of the mine. Placed in the suction pipe of 
the pump is a 300-H.P. Wainwright heater, with corrugated 
copper tubes. The water in the pump, at a temperature of 60°- 
70° F., passes through the heater tubes on its way to the pump 
suction valves. The air, on being exhausted from the high- 
pressure cylinder, at a pressure of 35 lbs., passes into the heater 
and through the spaces between the tubes. In this way, the 
temperature of the air is raised practically to that of the water 
and, after expanding again in the low-pressure cyhnder, is 
exhausted without freezing. Should the sump water be foul, 
the heater tubes must be cleaned from time to time; otherwise 
the coating of sediment materially reduces their conductivity. 
Still better results would be obtained from such an installation 
by employing a fly-wheel pump with a shorter cutoff. The 
lower temperature could then be met by water-jacketing both 
cylinders, the jackets being supplied with water by a small pipe 



408 COMPRESSED AIR PLANT 

from the pump column. Though the quantity of heat thus 
restored to the expanded air is far smaller than that which would 
be derived from a fuel-burning reheater, this simple device is 
convenient and satisfactory for underground service. 

By employing reheating in connection with properly designed 
and operated air-driven compound pumps, eihciencies of 40-50% 
may be realized. With 3-cylinder pumps, furnished with inter- 
mediate heaters, the efficiencies are still higher, reaching even 
70%. Reference has already been made to the economic 
advantages of using the Cummings system of high-pressure 
transmission for operating compressed-air pumps. 



CHAPTER XXV 
PUMPING BY THE DIRECT ACTION OF COMPRESSED AIR 

In the apparatus included under this heading, the compressed 
air acts directly on the water or other fluid to be pumped; there 
is no piston, cylinder, nor other moving part except the valves. 
The limiting depth or head at which these pumps will work 
depends on the gage pressure and mode of using the air. When 
operated under proper conditions and with expansive use of the 
compressed air, they compare favorably in first cost and efficiency 
with ordinary simple-cylinder piston pumps, and their main- 
tenance cost is very low.* 

There are three classes: 

1. Pneumatic-displacement pumps, using compressed air with 
or without expansion. 

2. The " Return- Air System." 

3. " Air-lift " pumps, always working expansively. 
Pneumatic-Displacement Pumps are of several kinds. In 

the type form, the compressed air acts directly upon the surface 
of the water contained in a submerged chamber or tank, valves 
being provided for controlling admission of air and water. The 
water is displaced by the air and is discharged from the tank 
through a pipe. There may be one or two tanks, the discharge 
pipe in the latter case being common to both. With one tank, 
the flow from the pipe is intermittent; wdth two, practically 
constant, the pair of tanks then having the same relation to each 
other as the chambers of the pulsometer pump. The working 
head is that which corresponds to the air pressure employed. 

* A valuable text-book, " Pumping by Compressed Air," by E. M. Ivens, 
member of Amer. Soc. Mech. Engs., was published in 1914 by John Wiley & Sons. 
To this book the reader is referred for a fuller discussion of the theory and practice 
of the subject than is appropriate here. 

409 



410 



COMPRESSED AIR PLANT 



Besides being useful for all low-head service, as pumping from 
wells or into tanks, pneumatic-displacement pumps, having 
practically no moving parts, are distinctly advantageous for 
pumping acids, chemical solutions, etc., which would rapidly 
destroy a piston pump. 

The double-chamber Merrill pneumatic pump is shown dia- 
graromatically by Fig. 218. Compressed air enters through an 



^Discharge 




Water Line 



Fig. 218. — Merrill Pneumatic-Displacement Pump. 



automatic valve, which opens connection alternately with the 

water chambers. The air pressure required depends on the 

height of Hft. Since the pressure per sq. in, of a column of 

water is 0.434 lb. per ft. of head, the height to which a given 

air pressure wiU raise water is equal to the gage pressure 

divided by 0.434; thus, air at 80 lbs. will pump to a height of 

80 

=184 ft. To cover friction, leakage, absorption of air by 

.434 

the water, and to provide dynamic head for overcoming inertia 

and securing a proper speed of discharge, an added air pressure 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 411 

is necessary. In terms of volume, i cu.ft. of water will be 
displaced per cu.ft. of compressed air. One cu.ft. of air at 

80 lbs. = = 6.33 cu.ft. free air. To this should be 

15 ^^ 

added for losses, etc., say 20%, making a total of 7.6 cu.ft. 

free air per cu.ft. of water. Taking i gal. of water 6qual to 

0.134 cu.ft., the work done per cu.ft. of compressed air, against a 

t8a 
head of 184 ft, will be: -=i8o ft.-gals. = i,sos ft.-lbs. 

In some cases more than 20% should be allowed. The actual 
work done in compressing i cu.ft. of air to 80 lbs. gage, by a 
single-stage compressor (Table V, Chap. X) is 0.183 H.P., or 
6,039 ft.-lbs. ; hence, the efhciency of the pump, with the above 
allowance for losses, is nearly 25%, which compares favorably 
with the efficiencies of simple direct-acting pumps. 

The displacement pump in its usual form exhausts at each 
stroke a tankful of air practically at gage pressure. By employing 
a series of these pumps in a shaft, and using the air expansively, 
the possible height of lift with a given initial pressure and the 
total efficiency, will greatly exceed that shown above.* This 
can be done by a suitable valve control, by which the air is 
expanded from the lowermost tank to the one next above, and 
so on, for smaller and smaller Hfts toward the top of the series. 
When the last tank is discharged, the whole system is occupied 
by expanded air, at a pressure of 2 or 3 lbs., which is then 
exhausted into the atmosphere. Air is admitted by the valve 
at intervals into the lowest tank, and the working of the system 
proceeds automatically. At 80 lbs. air pressure, water can 
thus be raised to a height of about 330 ft., instead of 184 ft., 
as in the preceding example, and at an efficiency of about 40%. 

Another displacement pump is the Latta-Martin (Fig. 219), 
designed chiefly for raising large volumes of water under low 
heads, though it may be constructed for any desired air pressure 
and head.t A pair of submerged tanks take water through 

* This series system of tanks has been proposed by E. A. Rix, Trans. Tech. 
Soc. of the Pacific Coast, Aug. 3, 1900, p. 187. 

t See also Compressed Air Magazine, Jan. 1907, p. 4332. 



412 



COMPRESSED AIR PLANT 



large disk valves in the bottom. The valve mechanism, admit- 
ting air alternately into each tank, comprises a main and auxihary 
valve, each thrown by a piston valve, like those of single- cylinder 
steam pumps. The valves are thrown by the oscillations of a 
pair of levers, from each of which is suspended a bucket filled 
with water and hanging in a housing within the main tank. 
When the pump is in operation, the bucket housings are alter- 




FiG. 219. — Latta-Martin Displacement Pump. 



nately filled and emptied, so that the difference in effective 
weight of the buckets causes them to rise and fall. 

The Halsey pneumatic pump (Pneumatic Engineering Co.) 
has a single, submerged tank, with a simple, automatic valve- 
motion, operated by a float. 

If a displacement pump be required to work in acid water, 
as in mines containing sulphide ore, the tanks may be lined 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 41^ 

with concrete and the other parts made of bronze; or the tanks 
may be replaced by excavations in the rock, adjacent to the 
shaft and Hned with concrete or asphalt. 

Return-Air Displacement System (Ingersoll-Rand Co.) uses 
compressed air expansively. Two tanks, either submerged or 
within suction distance of the sump, are connected by pipes 
with the compressor. If not submerged the water enters by 
siphon action. Compressed air admitted to one tank forces 
out the water (or other fluid) through a check valve to the pipe- 
line. Meanwhile, the compressor draws air from the other 
tank. The charge of air is so regulated that, when one tank 
is empty, the other is full; then a switch valve reverses the 
action of the tanks. 

Referring to the diagram (Fig. 220), A and B are the tanks, 
with their air pipe-lines A^, B^; C^ and C^ are discharge check 
valves, preventing return of the ejected fluid; D^, D^ are check 
valves, preventing discharge through the inlets; E is the discharge 
pipe of both tanks; F is the automatic switch controlling the 
pumping cycle; / is an automatic compensating valve, which 
keeps the system supplied with air, taking care of loss and 
leakage. Instead of being exhausted into the atmosphere at 
each stroke, the compressed air after doing its work is con- 
ducted back to the compressor intake and expands behind its 
piston. Hence, the system is a closed one, the same air being 
used over and over, as in the Cummings return-air compressor 
plant (Chap. XVII). 

In starting, after the water in one of the tanks has been 
expelled, the switch reverses and connects this tank with the 
compressor intake. Then, while the second tank is being dis- 
charged, the air exhausted from the first returns to the com- 
pressor and, acting expansively upon the intake side of the 
piston, reduces by so much the power required to drive the com- 
pressor. When the pressure in the first tank has fallen suf- 
ficiently (by being in communication with the compressor 
intake), it will again fill with water. Thus, the compressor 
transfers the same body of air from one tank to the other. Valve 
/ is set to open during the suction period, at a negative pressure 



414 



COMPRESSED AIR PLANT 



a little greater than that required to draw water into the tanks. 
The s\vdtch-valve is operated automatically, by a de\dce acting 
at the intervals required to complete a cycle in both tanks, or by 
an electric make-and-break mechanism, controlled by a pressure 




Tripping Cylinder 
Actuating Cylinder 



-E 



—Water 




Fig. 2 20.— " Return-.\ir " Displacement System (Ingersoll-Rand Co.) 

gage on the air intake. In the first case, a piston valve is 
operated by a small air cyHnder, compressed air being admitted 
alternately to each side of the piston in the latter through an 
auxihary valve. The volume of air for a given size of tank 
may be determined in terms of revolutions of the compressor. 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 415 

This pumping system has an efficiency of 55-60%, and 
requires Httle attention during operation. For further details, 
see paper by Prof. Elmo G. Harris, Trans. Amer. Soc. C. E., 
Vol. 54. 

Table XLVI 

Return- AIR System. Size of Compressor, Piping, Etc., 
FOR Different Heads, Based on ioo Gal. Water per 
Min. Sizes for other Quantities of Water are 
Proportional 



Lift, Ft. 


Capacity of 
Compressor, 

Cu.ft. per 
Min. Piston 

Displace- 
ment for 100 

Gals, per 
Min. 


Max. I. H.P. 

of Air 

Cylinder. 


Max. I.H.P. 
of Steam 
Cylinder. 


Aver.. H.P. 
of Steam 
Cylinders. 


Area of Air 

Pipe, Sq. in., 

for Each 

100 Gals. 

Capacity. 


Area of 

Water Pipe, 

Sq. in. for 

Each 100 

Gals. 
Capacity. 


SO 


39 84 


2.74 


3.22 


2.80 


.96 


7.70 


60 


42 


78 


3.28 


3-85 


3 


37 


1.03 


8.25 


70 


AS 


30 


3.85 


4-53 


3 


93 


1.09 


8.73 


80 


47 


70 


4-45 


5.22 


4 


49 


1. 14 


9. 12 


90 


49 


80 


S-03 


5-91 


5 


05 


I. 20 


9.60 


100 


51 


84 


5-67 


6.67 


S 


61 


1-25 


10.00 


120 


SS 


44 


6.96 


8.18 


6 


73 


1-33 


10.60 


ISO 


S9 


94 


9.00 


10.60 


8 


41 


1.44 


1 1 . 50 


170 


62 


64 


10 42 


12.25 


9 


54 


I SO 


12.00 


200 


66 


12 


12. 76 


15.00 


II 


22 


i-S8 


12.65 


250 


71 


TO 


16.46 


1935 


14 


02 


I. 71 


13.70 


300 


75 


06 


20.45 


24.00 


16 


83 


1.80 


14.40 



This table assumes that the tanks are fully submerged. 

Air-Lift Pump. This is a revival of an old principle. Since 
1888, when Dr. Julius Pohle proposed its application for pump- 
ing, the air-lift has become increasingly important. Besides 
its use for raising water from deep wells, it is applicable to a 
limited extent to pumping in shafts, and is well adapted for 
elevating finely divided pulpy material mixed with water, as 
the slimes and sands of cyanide and concentration mills. 

The pump consists essentially of two pipes; a large column 
or delivery pipe and a relatively small air pipe, leading from the 
compressor receiver (Fig. 221). The delivery pipe, open at both 



416 



COMPRESSED AIR PLANT 



ends, is submerged to a depth proportionate to, but always 
greater than, the height to which the water is to be raised. 

The compressed-air pipe 
passes down to a point near 
the bottom, and admits air 
to the lower end or foot- 
piece of the delivery pipe. 
( Modihcations of this ar- 
rangement are noted here- 
after.) 

In some respects the 
operation of the air-hft 
pump is the reverse in 
principle of the method of 
compressing air by the 
direct action of falling 
water (Chap. XV). If the 
discharge pipe be of very 
small diameter, the com- 
pressed air entering the 
bottom tends to form 
piston-like layers, which 
rise rapidly, alternating 
with masses of water (as 
is shown by experimenting 
with glass tubes). If the 
discharge pipe is of large 
diameter, the air is best 
admitted through a series 
of ports or nozzles, result- 
ing in a more complete 
dissemination of the air 
through the mass of rising 
water. The water is raised 
chiefly by the aeration of 
the column of water, which causes a reduction in its specific 
gra\ity; added to this is the expansive force and \is vi\a of the 







Fig. 221. — Diagram of Pohle Air-I.ift Pump. 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 417 

compressed air. Before the air is turned on, the water stands 
at the same level inside and outside of the delivery pipe. On 
entering the foot-piece, the air is under a pressure due to the 
weight of the rising column of water. As the bubbles of air 
rise, they expand with the decrease in head, and, on reaching 
the point of discharge, the tension of the air is reduced practically 
to atmospheric pressure. The initial air pressure required 
depends on the pressure due to head, measured from the point 
at which the air enters the delivery pipe to the surface of the 
water. At too high a pressure loss of work ensues at the com- 
pressor. Should the delivery pipe be too deeply submerged, 
in proportion to the net height of lift, an uneconomically high 
pressure will be required to force the air into the foot-piece; 
and, with insufficient submergence, more air will be necessary 
to produce the velocity of delivery. 
Referring to Fig. 221, let: 

h = depth of the delivery-pipe foot-piece below the normal water- 
level, before pumping begins, or when the water is at 
rest; 

/?2 = height of water-level when the pump is in operation ; 

H = height of the column of mixed air and water, measured from 
the air inlet to point of discharge;. 

L = net height of lift = H — 7^2. 

The compressed air enters the foot-piece at a pressure P', 
corresponding to the head h2\ or, /z2X pressure per foot of 
hydraulic head = 0.434 h2. Assuming that the water rises in 
piston-like masses (as would be the case with a single air nozzle 
and a delivery pipe of small diameter) , the sum of the length of 
these masses in the column H must theoretically be equal to the 
outside solid column of water 7/2 (the weight of the compressed 
air contained in the column being neglected). But, to over- 
come the frictional resistance and produce flow, the head 7^2 must 
be greater. Under ordinary working conditions, the net height of 
lift L is found to be from 0.5 7^2 to say 0.65 A2. Taking the second 

value and transposing: 7^2=— 7-; and by substituting iji the 



418 COMPRESSED AIR PLANT 

expression for the value of P^ as above: P' = 0.434—^ = 0.67 L. 

If, for example, L be 50 ft., P' = 33.5 lbs., and h2 = -^~ = y'j ft. 

Since the air in the column H is divided into small masses, 
surrounded by water, its expansion during the upward flow may 
be assumed to be isothermal. If P' be its initial pressure, the 

/P'\ 
mean pressure for the entire lift = PxNap.log( — I, P and P' 

being absolute pressures. In the above example, taking P as 
15 lbs., P' = 33.5 + 15 =48.5 lbs., whence, the mean pressure = 
17.5 lbs. gage. 

For starting the pump, the air pressure must be sufficient to 
overcome the normal static head hi, but, when the flow has begun, 
the pressure required falls to that corresponding to h2. Though 
this difference in pressure (^1-/^2) may be considerable, it is 
readily met by temporarily speeding up the compressor. To 
minimize fluctuations between hi and h2, the top of the well or 
sump should be extended laterally, in order to furnish a large 
horizontal area of water, the level of which would be but httle 
affected by stoppages or by variations in air pressure and delivery. 
The throttle valve in the air pipe may be regulated by a float on 
the surface of the water. Care should be taken in the design 
of the foot-piece and in properly proportioning the air pressure 
to the submergence and net lift. Otherwise, air may leak back 
into the sump or outside column of water; and, if this becomes 
aerated, much more power and a larger volume of air will be 
required to keep the pump in operation, thus reducing the effi- 
ciency. 

Since 1889 rnany experiments by competent engineers have 
been made on the air-lift pump. Among the first were those of 
B. M. Randall and H. C. Behr, on a 60-ft. well, with a stage 
compressor. A summary of these tests is given by E. A. Rix, in 
the Transactions of the Technical Society of the Pacific Coast, Aug. 
3d, 1900, p. 206. In 1894 a series of tests were made at De Kalb, 
111.,* and in 1893 ^^^ ^^9^ ^^ ^^^^ pumps at Rockford, 111. f 
* Eng. News, July 12, 1894. f Eng. News, March 4, 1897. 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 419 

The last-named were carefully carried out and the results 
compared in tabulated form. The heights of lift above water- 
level were 66.5, 90, and 91.5 ft., the air pressure being 76 lbs. 
gage and the submersion 225 ft. Both air pressure and depth 
of submersion appear to have been unnecessarily great. With a 
compressor of 124 H.P., the net work done was 24 H.P., or an 
efficiency of about 20%. With 600 cu.ft. free air per min., 
200 cu.ft. of water were puniped, or 3 air to i water. The 
sizes of piping used were: deHvery pipes, 4 ins., 5 ins., and 6i ins., 
with air pipes from ii-2| ins. In several of these tests the air 
pipe terminated in a f-in. nozzle. The plan was also tried of 
closing the lower end of the air pipe and discharging the air 
through slot-shaped perforations in the sides near the bottom; 
but the results were inferior to those obtained from the single- 
nozzle opening. Possibly better work would have been done by 
some different arrangement or size of slots; for large pipes and 
volumes of water the single nozzle has not been satisfactory. 

E. E. Johnson gives a table for computing the performance 
of the air-lift pump, including consumption of power and theo- 
retical and total efficiencies for different heights of lift,* from 
which Table XL VII is abstracted. 

These figures, which represent the work of well-proportioned 
plants, as to depth of submergence and air pressure, show that 
the efficiency falls off rapidly as the air pressure and height of 
lift increase. Under normal conditions and with small Hfts, 
efficiencies of 30 to 35% are readily obtainable, and may rise 
to 45 or 50% with proper air pressures and ratios of submerg- 
ence to height of lift. 

Volume of Air for Air-Lifts may be computed by the follow- 
ing formula, which closely approximates average practice, f 

in which: Va = vol. free air for raising i gal. water; ^ = total 
lift, ft.; H = submergence when running, ft.; C = constant. 

* Eng. News, April 22, 1897, 

t Ingersoll-Rand Co., suggested in part by E. A. Rix. 



420 



COMPRESSED AIR PLANT 



Values of " C '' with proper submergence: 

Lift in Ft. (//) Constant 

lo- 60 ft 245 

61-200 '' . 233 

201-500 '' 216 

SOI-650 '' 185 

651-750 " 156 

Table XL VII 





Theoretical Horse-power. 


EPFICIENC-i 


• OF Air 


-LIFT. 






Lift. 


• .^^ ft 



To Del 
Air 


iver I Cu.ft 
per Minute 


. of 


T 

Theoretical. 


otal Efficiency 
Tom Power 
Applied to 
Water Del'd. 


3 

V. 

.-J 




'a 
H 
<u 

g 


«3 

00 

■!-> 

6 




a 

< 


« 
E 


§0.2 c 

§S 1 


CS 4) 


6 ft 

HO 




<U y C 

iJ^ft 

^<E 

c 


5 


II 54 


.02185 


02514 


.02572 


0263 


.87 


848 


-83 


-623 


•497 


10 


23.09 


•04363 


05586 


-05992 


064 


.78 


728 


.684 


-546 




41 


15 


34 63 


• 06545 


09105 


.0962 


1015 


.72 


687 


.648 


•515 




•389 


20 


46. 20 


.08727 


12994 


•I39I 


1483 


•675 


627 


59 


•47 




354 


25 


57-75 


. 109 


I7I9I 


.1897 


2004 


-635 


575 


-545 


432 




327 


30 


69.31 


- 13091 


21678 


.2370 


2573 


.603 


548 


• 508 


412 




305 


35 


80.86 


•1527 


26445 


• 2915 


3187 


-577 


52 


478 


39 




287 


40 


92.41 


-17454 


31375 


•3489 


3842 


-557 


502 


455 


376 




273 


45 


103.90 


.1963 


36368 


-4085 


4535 


■540 


482 


433 


362 




260 


50 


115-50 


.21818 


41848 


-4722 


5261 


-522 


464 


415 


348 




249 


55 


127.00 


24 


47II2 


-5366 


6023 


-51 


447 


40 


336 




24 


60 


138.60 


.26181 


52855 


.6051 


6818 


•495 


432 


384 


324 




231 


65 


150.10 


.2836 


58612 


-6734 


7608 


-483 


422 


372 


316 




22s 


70 


161. 70 


30545 


64812 


.748 


8483 


• 471 


408 


36 


307 




216 


75 


173 30 


-3273 


70952 


-823 


9380 


.462 


398 


35 


299 




210 


80 


184.80 


3491 


76843 


.898 I 


0291 


•455 


39 


343 


292 




206 


85 


196.30 


37 


83039 


-976 I 


1231 


-446 


i^ 


331 


285 




198 


90 


207 . 90 


3927 


89444 


I 055 I 


2176 


-439 


373 


324 


28 




194 


95 


219.40 


-4145 


96164 


I 137 I 


3148 


•431 


268 


315 


276 




189 


100 


230.90 


. 43636 I 


0243 


1-247 I 


4171 


-428 


352 


308 


264 




185 


no 


254.10 


.48 I 


162 


1-394 I 


626 


413 


346 


296 


26 




177 


120 


277.20 


•5236 I 


301 


I 571 I 


841 


.402 


2>ii 


285 • 


25 




171 


130 


300.40 


-5675 I 


443 


1-755 2 


068 


394 


324 


275 ■ 


243 




165 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 421 

Air Pressure. Starting pressure is equal to the depth of 
foot-piece in the well, less the water head in ft. X 0.434. 
Working pressure is equal to the depth of foot-piece, less the 
pumping head X 0.434+ friction in air pipe +2 lbs. back pres- 
sure in the nozzle or foot-piece. 

Ratio of Lift to Submergence. The following figures repre- 
sent average practice.* 

For lifts up to 50 ft., 70-66% submergence. 

50-100 " 66-55% 

100-200 " S5-5o% 

200-300 " 50-43% 

" 300-400 '' 43-40% 

" 400-500 '' 40-33% 

Tests. In several tests at Wandsworth, England, on a 
modified Pohle air-lift, with a delivery pipe of increasing diam- 
eter toward the top, the total height of the delivery pipe was 
580 ft., of which 324 ft. were submerged, the net lift thus being 
256 ft. In this case, the distance }ii — h2 (Fig. 221) was 69 ft, 
air pressure, 135 lbs., ratio of volume of free air used to water 
discharged, 5.8 and 5.6 .1; total efficiency, 36%. This indi- 
cates an advantage in using a tapering column pipe. 

The following results of a test on a 300-ft. well show the low 
efficiency of high lifts :f 

Elevation of discharge above mouth of well 85 ft. 

Depth to water-level during operation of pump 44 ft. 

Net lift, water-level to point of discharge 129 ft. 

Submergence of deh very pipe 248 ft. 

Air admitted to delivery pipe 5 ft. above inlet end. 

Diameter of delivery pipe. 3.5 ins. 

Diameter of air pipe 1.25 ins. 

Volume of water delivered per minute 82 . 5 gals. 

Volume of free air used per minute 81.8 cu.ft. 

Gage pressure of air 107 lbs. 

Consumption of free air per cu.ft of water 7-44 cu.ft. 

Horse-power consumed by compressor 12. i 

Total efficiency 22 . 3% 

* Recommended by Sulhvan Machinery Co. 

t G. C. H. Friedrich, Trans. Ohio Soc. of Mech., Elec, and Steam Engrs., igo6. 
For data on a number of other air-lifts, see Peele's " Mining Engs. Handbook," 
igi8, Section 15, Art. 23. 



422 



COMPRESSED AIR PLANT 



Though the question of relative sizes of air and delivery 
pipes has not yet been satisfactorily answered, it is probable 
that ratios of diameter from i : 2 up to i : 2| or 3 will be found 
suitable. The absolute diameters are determined on the basis 
of frictional loss caused by the flow in the pipes. A water 
velocity of 250-300 ft. per min. may be assigned for the delivery 
pipe. For the friction losses in air pipes, see Chap. XVI. When 
the water is deHvered at a distance from the pump, the added 
frictional resistance must be determined, and the air pressure and 
submergence correspondingly increased (see a paper by H. T. 
Abrams, in Compressed Air Magazine, Aug., 1906, p. 4135). 

Table XL VIII gives some calculated values for air-lifts. 

Table XLVIII 



Lift, Ft. 


Volume of Air 

per Cu.ft. of 

Water. 


Submergence, at 
60% of Total 
Height of Delivery 
Pipe. 


Air Pressure. 


H.P. per Gal. 
Water per Min. 


25 


2 


38 


17 


0.0184 


50 


3 


75 


2>1> 


0.0426 


75 


4-5 


113 


49 


0.0828 


100 


6 


150 


65 


0. 1320 


125 


7-5 


188 


82 


0.1910 


150 


9 


225 


98 


0.2544 


175 


10.5 


263 


115 


0.3150 


200 


12 


300 


130 


0.3808 



Foot-Pieces for Air-Lifts are varied in design (Fig. 222). 
At A is shown the original Pohle side inlet, now rarely used. 
The annular foot-piece B (Ingersoll-Rand Co.) is a modification 
of A , admitting air all around the periphery of the delivery pipe, 
without contracting the discharge area. It is furnished for 
capacities of 20-325 gals, per min. In the Saunders foot-piece 
C, for a cased well, the air passes down through the annular 
space between the casing and delivery pipes. Another form of 
the Saunders foot-piece admits air to a hollow base plate studded 
with a group of vertical i-in. pipes, 18 ins. long. The mass of 
water is thus split up and the air mixed with it at the point of 
entering the delivery pipe. Design Z) is a variation of C. 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 423 

The design shown at E (Fig. 222), made by the SulKvan 
Machinery Co., and both of those in Fig. 223 (Ingersoll-Rand 



..>-.-- ■^y3?^ y ^/^///////^//;z^ 








^ 



^Q 



^ 



X 



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X 



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I 



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s 



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to 
c 



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I 

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Co.), admit air to the deHvery pipe through numerous small 
holes, to secure a thorough intermixture of the air and water. 
The Sullivan foot-piece is a bronze casting, and is furnished for 



424 



COMPRESSED AIR PLANT 







u 



capacities of 5-1,500 gals per min. In those shown in Fig. 223 
the outer casing is steel, the inner tube of brass; their capacities 
are 15-2,000 gals, per min. To give an accelerated velocity to 
the mixed air and water on entering the delivery pipe, the three 

designs last mentioned have a Venturi 
contraction throat above the foot- 
piece.* 

Elaborate tests on the air-lift pump 
were made in 1907 by Messrs. Hender- 
son and Wilson at the two 200-stamp 
mills of the Angelo and Cason mines, of 
the East Rand Proprietary Mines, 
Limited, South Africa. | At these mills 
slimes and sands are raised to the 
settling tanks by air-hfts instead of the 
usual taihngs-pumps and wheels. The 
delivery pipes used in the 19 recorded 
tests were of two kinds, viz., 10- to 
i6-in. pipes of constant diameter, and 
several pipes increasing in diameter 
from 12 and 14 ins. at the bottom to 
14 and 16 ins. at the top. As a uniform 
taper was impracticable, the latter 
pipes, for a length of 35 ft. above the 
air inlet, were lined with i in. of wood, 
which served also to protect the metal 
from the scouring action of the mixed sands or slimes and 
water. 

The foot-piece used in the earher tests was flared out and 
closed at the bottom, the water and pulp being admitted through 
4 large ports, 2 J ft. below the air inlet and having a combined 
area of about 200 sq. ins. The air inlet was a single opening, 4 
ins. diameter. For the later tests, the foot-piece was open at 
the bottom and flared out to double the diameter of the column 



0: 



It 



Fig. 223. — " VA " and 

" VC " Foot-pieces 

(Ingersoll-Rand Co.) 



* For further details of air-lift apparatus see the catalogues of the Sullivan 
Machinery Co., and Ingersoll-Rand Co., which contain much useful information, 
t For full details see The Engineer (London), Jan. 10, 1908, p. 26. 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 425 

pipe, to increase gradually the velocity of inflow (Fig. 224). 
A ring of 12 holes, i in. square, admitted the air. 

The modified foot-piece was supported on timbers, so that 
the entire bottom was open for the free admission of the material 
to be pumped. The column pipe was of steel tubing, expanded 

T 




Fig. 224. — Foot-Piece for Air-Lift Pump, for Raising Mill Tailings and Slimes. 



into cast-iron flanges, and lined in the lower part with wood. 
This design gave materially higher efficiencies than the one first 
used. Table XLIX, though presenting the details of only 4 
of the 19 tests made, indicates the general results obtained. 
These details show that the air-lift, when properly designed for 
stated conditions, can compete successfully with the tailings 



426 



COMPRESSED AIR PLANT 



wheel, in common use in the district, and that it is superior to 
the taiUngs pump. 

Table XLIX 





Test. 


I 


2 


1 
3 i 4 


.2 

c 
o 
U 


Number and size of delivery 
pipes 

Submersion in ft 

Lift in ft 

Ratio of submersion to lift. . . 

Gage pressure of air, lbs 

Kind of foot-piece 

Throat diameter of foot-piece. 


Two lo-in. 

32.75 

32.5 

I . 009-1 

15 
Original 
10 ins. 


Two lo-in. 

35-75 

29 -5 

I . 21-1 

16 

Original 

10 ins. 


One i6-in., 
decreasing 
to 14 ins. 

37-75 

27.5 

I. 372-1 

17 

Modified 

13^ ins. 


One 14-in., 
decreasing 
to 12 ins. 

48.85 
27.09 

I -77-1 

22 

Modified 

112 ins. 


c 

O 
i-i 


Free air, cu.ft. per min 

Free air, per cu. ft. of slimes. 

Cu.ft of slimes per min 

Throat velocity, cu.ft. per sec . 
Theoretical H.P. in pulp raised 
H.P. per cu.ft. free air com- 
pressed 

Air horse-power 


2256 

7.27 
310 
4.7 
19-3 

.048 

ior,.72 


1279 
4.06 

315 

4-8 
17.8 

.050 
64.74 


746.48 

2.74 
290 

4-85 
15-23 

-053 
42 . 21 


846 

2.64 
320 

7-39 
16.6 

.064 
54-14 




Efficiencv, per cent 


17.7 


27-5 


37.15 


30.55 







The conditions were modified in the successive tests, as to 
the ratio of submersion to Uft, diameter of deUvery pipe, and air 
pressure. As a basis for computing the horse-power represented 
by the mixture of water and pulp raised, the weight of the sHmes 
was determined to be 63.3 lbs., and of the sands, 64.56 lbs., 
per cu.ft. Thus, for the sands, the horse-power was: 



(Quantity of sands -f water) X 64. 56 X ft. lift 

33,000 



= .ooi956xQXft. lift. 



The term " sands " refers to the mixture of water and ore as 
crushed by the stamps, from which the '' slimes " have been 
separated in the milhng process.* 

* For more data on air-lifts for mill pulp, see Compressed Air Magazine, May, 
1914. 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 427 

LanselPs Air-Lift is a modification of the air-lift, applied by 
Mr. George Lansell to pumping from a deep mine shaft in the 
Bendigo district, Victoria, Australia. A series of lifts were used 
for a depth of 1,385 ft. Fig. 225 shows diagrammatically two 
of the lifts. 

The compressed air is conveyed from the receiver in a pipe A , 
running down the shaft. The water is conducted from a tank 
through a pipe D, which first passes down the shaft a certain 
distance, depending upon the height to which the water is to be 
raised, and is then connected with an enlarged section of pipe 
E, at the foot of the delivery pipe B. Thus, the piping for each 
lift forms an inverted siphon. At the lowest point of the siphon 
a short branch pipe C enters from the air main A, the end of 
this branch being directed upward into the foot-piece E. Before 
the compressed air is turned on the water stands at the same 
level in pipes D and B. The effect of this arrangement is like 
that produced by submerging the lower part of the ^delivery 
pipe, as in the ordinary air-lift. Check valves are placed, as 
shown, in pipes D and C, to prevent air or water from passing 
back into the air pipe or into the tank. A throttle valve in 
pipe C regulates the supply of air. The relative heights of the 
various parts are variable, the dimensions shown on the sketch 
indicating substantially the proper depth of the inverted siphon 
below the tanks, and the corresponding height of lift; thus, 
from the tank at the 250-ft. level, the pipe D passes down the 
shaft 140 ft., to the foot of the delivery pipe which discharges 
at the surface. By a series of lifts the water may thus be 
raised from any desired depth. The air pressure is the same 
for all, this pressure being 60-80 lbs. per sq. in., or that which is 
ordinarily furnished for mine service. 

Air-Lifts at the Old Dominion Copper Mine, Arizona.* Ten 
inch air-lifts (Fig. 226) were installed from the 12 th to the loth 
level (200 ft.) and from the loth to the drainage tunnel (431 ft.). 
The air lines were 4-in. The submergence on the lift from the 
12th level was 177 ft., or 47%; capacity, 1,650 gals, per min.; 
air pressure, 80-90 lbs.; air consumption, 1,080 cu.ft. per 1,000 
* P. G. Beckett, Trans. A.I.M.E., Vol. 55 (1916), p. 53- 



428 



COMPRESSED AIR PL.\NT 



mmT-/: 




Fig. 225.— Diagram of LanseU's Air-Lift Pump for Mine Shafts. 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 429 



i Ping, wijb Holes 
Pressure Drilled in it - 




Fig. 226.— Air-Lift at the Old Dominion Mine, Globe, Ariz. {Trans. A.I.M.E , 

Vol. 55). 



430 



COMPRESSED AIR PLAXT 



^Rubber 1 i 




Fig. 227.— Two-Stage Air-Lift, Coolgardie, Western Australia. 



PUMPING BY DIRECT ACTION OF COMPRESSED AIR 431 

gals.; maximum efficiency, 36%. Submergence on the loth 
level lift was 188 ft., or 30.4%; capacity, 1,233 gals, per min.; 
air pressure, 90-100 lbs.; air consumption, 2,681 cu.ft. per 
1000 gals. ; maximum efficiency, about 30%. 

During a period of unwatering the i6th level of the mine, 
the columns of two direct-acting pumps were converted into 
air-lifts. 

Two-Stage Air-Lift at Burbanks Main-Lode mine, near 
Coolgardie, Western Australia, for raising and transporting mill 
sands (Fig. 227).* The height of lift for single-stage- would 
have required sinking a pit 39! ft. deep, for the necessary sub- 
mergence; for two-stage, the pit was only 15 J ft. Above the 
pit are two wooden head-boxes, with water levels 9 ft. apart. 
The wells for the pulp are of 6-in. cast-iron pipe; inside pipes 
are 3 in., and the air pipes f in. diameter. The air outlets of the 
latter are slots, at an angle of 45°, thus giving the air jets an 
upward direction. Supply of air is regulated by floats, so that 
a nearly constant level is maintained in each head-box, and no 
air is blown to waste. As the sand is heavy, the pulp mixture 
is 2 J water to i sand. Results obtained are as follows: free air 
per min., 47 cu. ft.; sand per min., 320 lbs.; water per min., 
800 lbs. ; sand per cu.ft. free air, 6.8 lbs. ; sand and water per 
cu.ft. free air, 23.83 lbs. 

Note. — In 1919 and 1920, several large air-lifts were designed and installed 
by the Supt., S. F. Shaw, for unwatering the Tiro General Mine, Charcas, Mexico.f 
The net lifts have ranged from r6o to 1080 ft., and the water pumped from 275 to 
900 gals, per minute; pumping efficiency, 20% to nearly 40%. 

As the work is still in progress (April, 1920), discussion of the operating results 
must be postponed, but it should be stated that these results show that the large 
percentages of submergence customary hitherto are not in reality necessary. At 
1,060 ft. lift, and submergence of only 30 ft. (2.8%) no gals, per minute were raised 
at an efficiency of 30.7%; and the last installation, raising 200 gals, 1,080 ft., with 
•213 ft. submergence (19.6%), is operating at about 30% efficiency, using 7.25 cu. ft. 
free air per gallon. 

* Jour. Chamber of Mines of Western Australia, Nov., 1910; abstract in 
Eng. b° Min. Jour., Apr. 8, 191 1, p. '706. 

t Trans Am. Inst. Min. Engs., Feb., 1920, also personal communications to the 
author. 



CHAPTER XX\T 
CO:\IPRESSED AIR HAULAGE 

For underground transport, compressed-air and electric 
locomotives divide between them a broad field of operation. 
Both are apphcable to mine ser^-ice of all kinds, for hauls of 
a few hundred feet to several miles. But for coal mines con- 
taining fire-damp, while compressed air is perfectly safe, electric 
locomotives must be adopted with caution. Although, by the 
improvements of recent years, much has been done to prevent 
serious sparking at the motor, some risk still exists; and, further- 
more, the possibility of strong sparking, accompanied by the 
momentary development of intense heat, from short circuiting 
or a ruptured conductor, can hardly be averted. 

Besides its safety for gassy, or dry, dusty collieries, or in 
dry and hea\'ily timbered workings, compressed air haulage 
has the follo^^'ing advantages: first, since the power is stored 
in the locomotive itself, the system has the maximum degree of 
flexibility; the locomotives can go wherever track is laid, far 
beyond the end of the supply-pipe fine. Electric locomotives, 
except those having storage batteries, are dependent upon 
wiring, which must accompany every foot of ad\ance.* For 
collieries compressed air may be used equally well for main-Hne 
haulage, and for gathering and distributing cars among the work- 
ing places; seco?id, compressed air costs Httle or nothing when 
not in actual use, and its full power or but a fraction of it is 
available at all times. In the intervals between hauls, no pov^er 
is w^asted, because, though the compressor may continue run- 
ning, it is engaged in storing up power in the pipe-Hne. A 
minor consideration is that the exhaust of the locomotive 

* " Cable-reel " or " gathering " electric locomotives are useful for very short 
distances only. 

432 



COMPRESSED AIR HAULAGE 433 

discharges fresh air into the workings, improving the ventila- 
tion of the mine. 

At most mines compressed air is employed only for under- 
ground transport, from the stopes or breasts to the hoisting 
shaft. In other cases, where the mine is worked through a 
tunnel or adit, trains are hauled direct to the breaker, tipple, 
or ore-bins, on the surface. Occasionally, as at the Homestake 
Mine, Lead, S. D., compressed-air locomotives are used for 
surface transport of ore, from the crusher houses at the shaft 
mouths to the different stamp mills, the object being chiefly 
to reduce the fire risk for the wooden structures, into and near 
which the haulage tracks pass. For the same reasons com- 
pressed-air haulage may be installed at lumber yards, or fac- 
tories containing inflammable buildings or materials. 

Compressed-air locomotives were probably first used in the 
works of the Plymouth Cordage Co., Plymouth, Mass., about the 
year 1873, and in Great Britain, for mine haulage, in 1878, but 
the early designs were not very successful. In the United States 
perhaps twenty compressed-air locomotives were built previous 
to 1898, but since then they have been applied for a great variety 
of service. In general terms, the plant consists of a compressor 
(usually three-stage), receiver, pipe-line, charging stations, with 
the necessary valves, and one or more locomotives. The loco- 
motive storage tanks are charged with a sufficient volume of 
high-pressure air for a round-trip run of the maximum length 
required, after which the locomotive returns to the nearest 
charging station for a fresh supply of air. 

Construction and Operation of the Locomotive. For mine 
service the locomotive generally has one or two storage tanks, 
which, with the cylinders, piping, and other appurtenances, 
are mounted on a frame carried by 4 or 6 driving wheels. 

Previous to 1908, practically all compressed-air locomotives 
were single-stage. Due to their greater efficiency and saving in 
compressed air, compound (two-stage) locomotives are now the 
rule, though single-stage engines are still built. Fig. 228 shows 
a recent design of a 4-wheel, twostage locomotive, made in 4 
sizes (Table L). 



434 



COMPRESSED AIR PLANT 



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COMPRESSED AIR HAULAGE 



435 



Small two-stage locomotives have the high-pressure cyHnder 
on one side, the low-pressure on the other. Larger sizes have 
four cyHnders, a high- and low-pressure, tandem, on each side. 
The initial air pressure is 250 lbs., which, by expansion in the 
high-pressure cylinder, is reduced to 50 lbs. (for ordinary con- 
ditions the cylinder volume ratio is i : 4) . The expanded air 
leaves the high-pressure cylinder at about 140° F. below atmos- 
pheric temperature (say 75°-8o° below zero F.). Reheating 
between the cylinders is therefore necessary. 




Fig. 228. — II. K. Porter Co.*s Locomotive, Class B-P-0 (see Table L.) 



Fig. 229 shows diagrammatically an efficient " atmospheric " 
reheater, consisting of a cylindrical shell filled with small brass 
or aluminum tubes, between which the exhaust air passes (see 
also Fig. 228, in which the reheater is shown in position). The 
tubes are open at both ends, the reheating medium being atmos- 
pheric air, drawn through the tubes by the ejector action of the 
exhaust from the low-pressure cylinder. As the volume of the 
high-pressure exhaust air is relatively small, its temperature 
is raised almost to that of the atmosphere. 

' A 6-wheel, single-stage locomotive, by the Baldwin Loco- 
motive Works, is shown in Fig. 230. Dimensions: gage, 3 ft.; 
cylinders, 11 by 14 ins.; 2 main tanks, 22 ft. 7 ins. and 20 ft. i in. 
by 34 ins. diameter, carrying a pressure of 800 lbs.; auxiliary 



436 



COMPRESSED AIR PLAXT 



tank pressure, 140 lbs.; driving wheels, 28 ins.; wheel-base, 
total, 6 ft. 6 ins.; total weight, 39,050 lbs., all on driving wheels. 
Another Baldwin locomotive, of the 4-wheel t}pe, with 9 by 
14-in. single-stage cylinders, 5-ft. 6-in. wheel-base, and weighing 
24,350 lbs., is sho^Ti in Fig. 231. These builders make other 



Pop Yalve 



Whistle 



TLroUle 




iiow-pressure Cylinder 



Fig. 229. — Diagram of Reheaterand Cylinders for H. K. Porter Co.'s Two-Stage 

Locomotiv^es (Table L). 



sizes, the smallest weighing 8,000 lbs., and having 5 J by lo-in. 
cyhnders; track gage, 36 ins.; tank pressure, 900 lbs., and 
working pressure 170 lbs. Working pressures of single-stage 
locomotives generally range from 140-180 lbs. Compressed- 
air mine locomotives are built also by the American Locomo- 
tive Co. 



COMPRESSED AIR HAULAGE 



437 



For track with sharp curves, the wheel-base must be short, 
say 4 ft. 6 ins. to 6 ft., for a 4- wheel engine. The height of the 
locomotive over all depends somewhat on the conditions existing 
in the mine as to thickness of vein, headroom of the haulageways, 
etc. ; it is rarely more than 5 or 6 ft., frequently less. The length 




Fig. 230. — Single-Stage Locomotive,. Baldwin Locomotive Works. 




Fig. 231. — Single-Stage Locomotive, Baldwin Locomotive Works. 

varies mainly according to the tank capacity required and the 
curvature of the gangways. It is usually 10-15 ft- for the 
smaller sizes, to 20 or 24 ft. for the larger. Widths, 3^-6 ft. 

Fig. 232 shows an H. K. Porter Co., 4-wheel, compound 
locomotive, made in 4 sizes, as detailed in Table LI, and Fig. 
233 a 6- wheel compound (Table LII). The hauling capacities 



438 



COMPRESSED AIR PLANT 




9 

a, 

I 



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442 COMPRESSED AIR PLANT 

in these tables are a maximum for the stated conditions; for 
satisfactory work train weights should be 50-90% of those 
tabulated. 

Fig. 234 shows a small locomotive of special design, with 
4 by 7-in. cylinders, used on track ,of 27-in. gage for hauling 
cars from underground loading chutes to the shaft stations. 
The operator's seat is detachable, so that the locomotive can be 
transferred from one level to another, on a cage with a 5-f t. plat- 
form. This design is suitable for general service in small metal 
mines, or gathering cars from working places in collieries, to 
make up trains on main haulageways. For mines having very 
steep local grades, the H. K. Porter Co. builds a locomotive 
like that in Fig. 228, but with a cable-reel on the rear. It 
operates like an ordinary locomotive on easy grades, and serves 
as a hoisting engine to gather cars from steep cross entries, 
stopes or breasts. The reel or drum, holding 300-1,000 ft. 
of wire rope, is driven by a pair of small auxiliary cylinders. 

Fig. 235 is a. diagram of the rear end of a Porter two-stage 
locomotive like that in Fig. 228. Additional features of the 
construction of compressed-air locomotives are given in 
Fig. 236. 

Construction Details. The tanks have dished or approxi- 
mately hemispherical ends, and are built of heavy boiler plate, 
with a tensile strength of 60,000-68,000 lbs. per sq. in.; the 
shells are 4-J in. thick, with i-15-in. heads. Ring seams are 
double riveted with lap joints; longitudinal seams have butt 
joints, with inside and outside welt strips. As the tanks are 
generally designed for working pressures of 700-900 lbs., the 
longitudinal seams have 6-8 rows of rivets, to make a joint of not 
less than 75% of the strength of the plate. Tanks are tested 
with a pressure about 30% greater than the working pressure, 
the factor of safety with plate of the usual quality being about 
3I; this is sufhcient, as there are no expansion and contraction 
strains, as in a boiler. For very high pressures, tanks of large 
diameter are unsafe, and are replaced by several smaller tanks 
(Fig. 237), or by a set of heavy seamless steel tubes, 8-9 ins. 
diameter. Tubes 9 ins. diameter by J§ in. thick will carry 



COMPRESSED AIR HAULAGE 



443 




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T^o face page ^44 



Compressed Air haulage 445 

pressUfes of 2,000-2,200 lbs. per sq. in. These high pressures 
are unnecessary for ordinary haulage service. 

From the main tank the air passes into a small auxiliary 
reservoir and thence to the cylinders. This reservoir is usually 
a section of wrought-iron pipe, 4-9 ins. diameter and 6-15 ft. 
long, laid alongside the main tank. The pressure in it is adjusted 
by an automatic reducing valve to the requirements of the 
engine — usually 140-180 lbs. for single-stage, and 225-250 lbs. 
for two-stage locomotives, depending on the size of cylinders 
and the power required. On the locomotives of the H. K. 
Porter Co., the reducing valve is a double-seated balanced 
valve, operated by a small piston. The air pressure in the 




Fig. 2j7. — H. K. Porter Co.'s Locomotive, Class C-5PS-O. 

reservoir acts on one side of the piston, tending to close the 
valve. This action is opposed by an external spring, adjusted 
to keep the valve open until normal working pressure is reached 
in the reservoir. Then the valve is closed by the air pressure, 
against the spring resistance. To provide for the case when the 
locomotive is using no air (as on a down grade or when at rest) , 
a single-seated supplementary valve is placed in the pipe between 
the reducing valve and the main tank. This valve is controlled 
by the throttle lever; being open when the throttle is open, 
otherwise closed by the air pressure. Having the two valves, 
leakage from main tanks to auxiliary reservoir is avoided and a 
close regulation secured. 

From the auxiliary reservoir the air passes to the cylinders 
through a balanced throttle valve. This maintains a constant 



446 



COMPRESSED AIR PLANT 



Ml 



^irt»- 



working pressure, suited to the needs of the locomotive, prevents 

waste of air Hkely to ensue if 
air at full tank pressure were 
admitted to the cylinders, and 
makes the locomotive more 
manageable. In starting a heavy 
load excessive slipping of the 
drivers is avoided, and with light 
loads the reducing valve may be 
regulated for any desired pressure. 
Toward the end of the trip, when 
the pressure in the main tanks 
falls to that in the auxiliary, the 
cylinders take air directly from 
the former, and the locomotive 
will continue to run as long as 
the pressure remains sufficient. 
For long hauls, or when the cross- 
sectional dimensions or sharp 
curves, or both, of the haulage- 
ways do not permit the use of 
tanks of great length or large 
diameter, a tender carrying a 
supplementary tank may be em- 
ployed (Figs. 238 and 246). 

For small, single-stage locomo- 
tives, the air is sometimes admit- 
ted to the cylinders throughout 
nearly full stroke, and conse- 
quently, as the exhaust is at high 
pressure, the efficiency is low. This 
practice is due to the tendency to 
use as small a locomotive as possi- 
ble, on account of the limited 
headroom and narrow, crooked 
gangways so common in mines. 

Better results are obtained by working with a cutoff, and increas- 






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P3 






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COMPRESSED AIR HAULAGE 447 

ing the size of the cyhnders and the weight on the drivers. In 
using air expansively, as can be done with properly proportioned 
cylinders, there should be no trouble from freezing of moisture. 
Although expansion produces a low cylinder temperature, yet, 
as the initial pressure is so much higher than is employed for 
other compressed-air machinery, the expanded air is relatively 
dry, and the force of the exhaust is sufficient to keep the ports 
clear of ice. To this end the ports should be large, straight, and 
short. The cyhnders are not lagged with non-conducting 
covering, as is so necessary for steam cylinders, to minimize 
condensation. By exposing their surface to the warm air 
of the mine, some heat is absorbed. Occasionally the exterior 
surface of the cylinders is cast with deep corrugations, to present 
a large area to the warm surrrounding air (Figs. 230, 231, and 
236). The cylinders are provided with slide valves; piston 
valves, like those used in steam locomotives, would leak more 
because of the dryness of the air. 

On account of the cold produced by reducing the pressure 
between the main tank and auxiliary reservoir, and to increase 
efficiency of operation, reheating is advantageous (though not 
essential) for single-stage locomotives. It is best done by heating 
the auxiliary reservoir. If steam be available in the mine, a little 
steam and hot water may be injected into the reservoir each 
time the locomotive is charged. Or, in mines free from fire- 
damp, a small reheater, burning oil or coke, may be carried on 
the locomotive. When the air is reheated water should always 
be kept in the small tank ; the moisture from it, passing with the 
air into the cylinders, assists in lubricating the valves and pis- 
tons. For two-stage locomotives, reheating is unnecessary, 
except between the cylinders (see Fig. 229 and accompanying 
description, p. 435). 

Pipe-Line and Charging Stations. The required capacity of 
the plant depends on the length of haul and size of locomotives, 
as influenced by the daily output, weight of trains, and track 
gradients. For short hauls, the pipe-line is som.etimes omitted, 
the locomotive returning to the compressor receiver to be 
recharged. Usually a pipe-hne is carried underground, with 



448 COMPRESSED AIR PL.\NT 

charging stations at one or more points, located according to 
the haulage distances and the capacity of the locomotive tanks. 
The innermost station, farthest from the compressor, must be 
at a point from which the locomotive can reach the end of its 
trip and return for recharging. For long hauls, hea\y traffic, 
or adverse gradients, several charging stations may be required. 
Inside receivers are unnecessary, unless the diameter of the 
pipe-hne is too small. The pipe-Hne itself acts as a storage 
reservoir, and should be of a diameter which, in proportion to its 
length, mil furnish a cubic capacity sufficient to charge the 
locomotive tank quickly and without excessive drop in pressure. 
That is, the pressure in the tank and pipe-line on equalizing 
should not fall much below the pressure which the locomotive 
is designed to carry. To this end, the volume of pipe-line 
storage should be at least three times the tank capacity of the 
locomotive. To determine the necessary pipe-line capacity, 
several variables must be harmonized, as follows:* 

V = storage volume required, cu.ft.; 

V = volume of locomotive tank, cu.ft.; 

P = pipe-line pressure, lbs. per sq. in. (usually 900-1,200 lbs.); 
p = desired pressure in locomotive tank, lbs. per sq. in. (700-900 

lbs.) ; 
p' — residua] pressure in locomotive tank, just before recharging, 

lbs. per sq. in. 

v(p-p') 



Then: Xil'- p) =v (p- p'), or Y = 



F-p 



For example, let P = 9oo lbs.. p = 7S^ ^bs., p' = i2^ lbs., and 
z;= 100 cu.ft., from which: 

1 00(750- 125) 

V = =416.6 cu.ft. 

900-750 

By transposition, the same formula serves for finding the 
pipe-fine pressure required for a given tank pressure. When 
there are several locomotives, it is rarely necessary to design the 
pipe-line for charging more than one at a time. If the volu- 

* H, K. Porter Co., " Modern Compressed-.\ir Locomotives," IQ16, 



COMPRESSED AIR HAULAGE 449 

metric capacity of the pipe-line be ample, the drop in gage 
pressure on charging is soon recovered by the compressor, which, 
except in plants operating a single locomotive, is kept in nearly 
constant operation. If more locomotives are added after the 
original installation of the system, the same pipe-line may still 
serve, provided the compressor is able to charge it to full pressure 
at shorter intervals. 

The piping, which is generally from 3-5 in., should be of the 
best material, lap-welded, and with sleeve joints made with the 
utmost care to prevent leakage. To stop leaks, the sleeves should 
have annular grooves (counter-bores) at each end, into which 
lead calking is driven if required. About every 300 ft. a pair of 
flanges should be placed in the line, and valves at suitable places, 
for convenience in making repairs or extensions. There should 
also be a valve between the compressor and pipe-line, so that 
the compressor can be inspected or repaired without wasting 
the air in the pipe. The pipe should not be buried, but carried 
along one side of the haulageway, either on the floor or on 
brackets, so that leaks will at once attract attention. An 
occasional bend in the pipe-line is advantageous in permitting 
free expansion and contraction, but bends should not be too 
numerous, as they involve more joints and greater possibility of 
leakage. Pipe-lines should be painted with some non-corrosive 
coating. 

Charging Stations. Fig. 239 shows one design, consisting 
of a heavy tee inserted in the air main, with a flexible connection 
for coupling to the locomotive. The connecting pipe comprises 
a vertical, rigid branch, with a ij-in. gate- valve, and a short 
horizontal pipe, with a union and a ball-and-socket joint. A 
further flexible connection, with 2 ball joints, serves for coupling 
to the locomotive tank. Thus, for charging, the locomotive 
need not be spotted accurately on its track, but has a foot or 
two leeway. When not in use, the flexible connection is swung 
aside. 

After coupling to the locomotive, the gate- valve is opened, 
whereupon the air pressure immediately forces together the 
parts of the ball-and-socket joints and makes a tight connec- 



450 



COMPRESSED AIR PLANT 



tion. When equilibrium is established between the pressures 
in the pipe-line and tank, the gate-valve is closed. To break 
the coupling, the compressed air remaining in the connect- 




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ing pipe, between the gate-valve and the locomotive check- 
valve, must first be released. This is done by opening a small 
'' bleeder valve," placed just above the gate- valve. The joints 
then loosen and are readily manipulated. The actual time 



COMPRESSED AIR HAULAGE 



45i 



Occupied in charging is usually about three-quarters of a minute, 
but, including stopping the locomotive and making the con- 
nection, iJ-2 mins. may be allowed. Charging may often 
be done while shifting cars and making up trains. Fig. 240 shows 
another form of charging station. 




T- Joint on 
Pipe Line 

Fig. 240. — Charging Station (H. K. Porter Co.). 

Calculation of Motive Power. Several factors must be 
known, viz., the tractive resistance per ton of the loaded cars 
on a level, resistances due to gradients and curves, Vvreight of 
empty and of loaded cars, and number of cars to be hauled per 
train. The values of these factors are readily ascertained, with 



452 COMPRESSED AIR PLANT 

exception of the resistances due to curvature of track and char- 
acter of roadbed. The former has been determined experimen- 
tally for ordinary surface railways, but mine track is apt to be 
roughly laid, with curves of varying and irregular radius, and the 
elevation of the outer rail improperly adjusted. 

The average tractive force required per ton depends on the 
condition of the track and roadbed, and the character and state 
of repair of the running gear of the cars. On level mine track 
the coefhcient of rolling friction is usually from 30-40 lbs. per 
ton, though it may be considerably higher on poorly laid or Hght 
track. With mine track in exceptionally good condition, this 
coefficient may be as low as 20 lbs. The grade resistance is 
20 lbs. per short ton, for each 1% of grade. The distribution 
of grades is often such that the maximum load is not the resistance 
of the loaded trains, which are usually hauled on slight down 
grades, but that of the return trains of empty cars on adverse 
gradients. For the most economical work, gradients should 
not exceed |-f of 1% in favor of the loaded trains. With 
ordinary track and rolling stock, and a grade of 5-6 ins. per 100 
ft., the coefficient of rolling friction is nearly the same for a 
loaded train hauled down as for an empty train of the same 
number of cars hauled up the grade. Heavier adverse grades 
are often necessary (2j%-3% or more), but they should be 
avoided, because the maximum tractive force of a locomotive 
falls off rapidly. On a 2j% adverse grade the locomotive can 
haul only about 4 times its own weight, even if the track be not 
slippery. Grades should be reduced on curves. Colliery cars 
carrying 2J-3I tons weigh 1,800 -2,300 lbs., while those used in 
metalliferous mines, for mechanical haulage, weigh 1,000 and 
2,000 lbs. Having ascertained the values of the different factors, 
the proper allowance of reserve power, in terms of volume and 
pressure of air, to cover indeterminate resistances due to local 
imperfections of track and rolling stock, is a matter of judgment 
and experience. 

With a given air pressure, the capacity of the locomotive 
tanks depends primarily on the length of round trip to be made 
with a single charge of air. When this distance is i-i^ miles, 



COMPRESSED AIR HAULAGE 453 

the tank capacity is generally from 50 to 150 cu.ft., according 
to the load; which, in turn, together with the track and grade 
resistances, governs the cylinder dimensions. Cylinders of 5 by 

10 ins. up to 9 by 14 ins. are common, the larger sizes being 
for heavy work in collieries or on surface. For runs exceeding 

1 1 miles, it is often desirable to increase the air pressure, rather 
than the tank capacity. Another plan is to provide a tender, 
carrying an auxiliary tank (Figs. 238 and 246). 

Having determined the total foot-pounds of work to be done 
with a single charge of air, on a run of the maximum length, 
specifications may be obtained from the builders for a loco- 
motive of suitable weight, gage, wheel-base, tank capacity, and 
cylinder dimensions. 

Compressors for Charging Locomotives are three- or four- 
stage. The air cylinders of the higher stages are single-acting. 
Fig. 241 shows a three-stage locomotive charger built by the 
Norwalk Iron Works Co., for pressures up to 1,000 or 1,200 lbs. 
The air passes from the low-pressure cylinder to the lower of 
the two intercoolers and, thence to the intermediate cylinder. 
From the latter the air goes through the vertical pipe to the 
upper intercooler, and thence through the inclined pipe to the 
high-pressure cylinder, from which the compressed air is deliv- 
ered to the receiver through the connection indicated under the 
outer end of the cylinder. 

The air end of a three-stage locomotive charger, by the 
Ingersoll-Rand Co., is shown in Fig. 242. The high-pressure 
intercooler is in the lower right-hand corner of the cut. Figs. 
243 and 244 illustrate a duplex, four-stage compressor; in Fig. 
243 are the intake and first intermediate cylinders, and in 
Fig. 244 the second intermediate and high-pressure cyHnders. 
Fig. 245 shows tlie entire compressor. 

The pistons of the high-pressure cylinders are sohd rams, 
with a series of packing rings. These, with the high-pressure 
valves, must be made with especial care, to prevent the serious 
effects of leakage of high-pressure air. Locomotive chargers 
are also built by the Sullivan Machinery Co. and others. 

When the mine is already provided with an ordinary low- 



454 



COMPRESSED AIR PL.\NT 



pressure air plant, for machine drills, etc., and which has some 
surplus capacity, a two-stage charging compressor may be 
installed, to take air from the low-pressure system and bring it 
up to the tension required for the locomotives. Some reduction 




S 
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O 



O 



in cost of plant may thus be effected, but care must be exercised 
in making such a combination, and it is not advisable. If the 
quantity of air produced by the low-pressure system should 
at times be insufficient to furnish the necessary excess for the 



COMPRESSED AIR HAULAGE 



455 




456 



COMPRESSED AIR PLANT 




FiGJ. 243 and 244. — IngersoU-Rand Four-Stage Locomotive Charger. 



COMPRESSED AIR HAULAGE 



457 



locomotive charger, the latter might have to compress from 
too low an initial pressure. There would be excessive develop- 
ment of heat which might raise the cylinder temperature to the 
flashing-point of the oil, thus causing an explosion (Chap. XIV). 




Fig. 245. — Ingersoll-Rand Four-Stage Compressor. 



Capacity of the Charging Compressor depends on the pipe- 
line pressure, number of locomotives to be operated, cubic con- 
tents of the locomotive tanks, pressure carried by the system, 
and the relation between charging periods. 
Let C = compressor capacity required, cu.ft. free air per min.; 

L = locomotive-tank capacity, cu.ft. free air per min. ; 

N = number of charges required in any given time, T. 

riL 

Hence the equation : C = -7^^ 

For example, if N = 3, L = 5,200 (corresponding to 100 cu. ft. 

of air at 750 lbs. gage pressure), and T = 6o minutes: 

3X5,200 
C = — — = 260 cu.ft. free air per min. 

When the locomotives are charged at approximately equal 
intervals of time, a single application of the above formula will 
be sufficient. Otherwise, calculations are required to determine 



458 COAIPRESSED AIR PLANT 

the maximum and minimum rates of consumption of air. For 
plants installed at an altitude above sea-level, allowance must 
be made for decreased output (Chap. XIII). 

Examples of Compressed-Air Haulage Plants. 

1. At the Buck Mountain Colliery, Penn., are two 8-ton H. 
K. Porter locomotives, each with 2 tanks, 15 and 17 ft. long, 
having a combined capacity of 130 cu.ft. of air at 550 lbs. pres- 
sure. Cylinders, 7 by 14 ins.; wheel-base, 5 ft. 3 ins.; height, 
5 ft. 2 ins.; length over all, 19 ft.; gage of track, 42 ins. A 
round trip of 5,100 ft. is made in 30-40 mins., or 2,500 ft. in 
12-15 rnins, with 12-car trains, on grades of i-4i%, averaging 
f of 1% in favor of the load. One locomotive delivers 150 cars 
per 10 hours, doing the work formerly done by 15 mules. Weight 
of cars, 3,400 lbs. empty, and 10,400 lbs. loaded. A three-stage 
Norwalk compressor suppHes 375 cu.ft. free air per min., at 700 
lbs. gage. Pipe-line, 4 ins. dianieter and 9,600 ft. long; storage 
capacity, 800 cu.ft. 

Average cost per ton-mile: 1.875 cents for the gross weight 
hauled, or 3.77 cents for net weight of coal. The. cost for mule 
haulage under the same conditions was formerly 3.94 and 7.92 
cents, respectively. 

The cost of this plant was as follows: 

Two locomotives $5,So5 

Air line: 9,647 ft. of 4-in. pipe $2,894 

Six charging stations 360 

Fittings and v^alves 382 

Labor cost for erection 998 

4,634 

Compressor $2,880 

Sundries and erection 246 

Compressor house 256 

Steam line to compressor 152 3,534 

Total cost $13,673 

2. Empire Mine, Grass Valley, Cal. Several small loco- 
motives, built by Edward A. Rix, are employed in the deep 
levels, for hauling trains of 5 cars, each carrying i ton. Maxi- 
mum distance per round trip, about 5,000 ft. Locomotive tank, 
36 ins. diameter by 48 ins. long; pressure, 500 lbs; dimensions 



COMPRESSED AIR HAULAGE 



459 




460 COMPRESSED AIR PLAXT 

of locomotive, 5 ft. long by 30 ins. wide by 52 ins. high; gage 
of track, 18 ins. One locomotive (Fig. 246) is operated by a 
pair of vertical engines, a chain and sprocket drive connecting 
the crank-shaft with the rear axle. There are 2 tandem tanks, 
one of them carried on a tender. A reheater (Primus kerosene 
burner) reheats the air after its pressure has been reduced in the 
auxihary reservoir. ^Ir. RLx has built 3 similar locomotives, 
but with a single, larger tank, for a 3-mile timnel, near San 
Francisco. They carry 1,000 lbs. tank pressure, the working 
pressure being 100 lbs.; each makes a 2-mile round trip, at 6-7 
miles per hour.* 

3. The Peerless CoUiery, Mvian, West \'a., operated for 
years several H. K. Porter locomotives, with 5 by lo-in. cylinders 
and weighing 10,000 lbs. Over-all dimensions: 10 ft. 5§ ins. 
long by 5 ft. 8 ins. wide by 4 ft. 5 ins. high. Four driving wheels, 
2^ ins. diameter; gage, 44 ins. Capacity of storage tank, 47 
cu.ft. ; pressure, 535 lbs.; charging time, 20 seconds; working 
pressure, 125 lbs. Pipe-Hne, 3 ins. diameter, with a capacity of 
242 cu.ft. Line pressure, 735 lbs. Trains consist of 6 cars, 
each weighing, loaded, 8,500 lbs. Grades range from level to 
2^%, generally in favor of the load. Curves from rooms to 
haulageways, 23 ft. radius, though the locomotives can work on 
curves of 15-ft. radius. Maximum round trip, 9,000 ft.; maxi- 
mum speed, 10-12 miles per hour. Cost of each locomotive, 
$1,800. 

4. The following data, on one of the plants of the Phila- 
delphia & Reading Coal &: Iron Co., were compiled by Mr. 
G. Clemens, a division engineer of the Company: 

a. Shaft level— i locomotive. 

Round trip, 5,200 ft.; grades 1% to 1% of 1%, all in favor 
of load; charging station at each end of run; gage of track, 44 
ins.; weight of cars, empty. 3.300 lbs., loaded, 8,800 lbs.; 15-38 
cars per trip; total output, 600 cars per 10 hours. Round- 
trip time, 12 min.; charging time, i min. A round trip and a 
half can be made with one charging. 

b. Slope level — i locomotive. 

* Compressed Air Magazine, Feb., 1908, p. 4747. 



COMPRESSED AIR HAULAGE 46 1 

Length of haul, 3,200 ft., of which 700 ft. is on an adverse 
grade of 4x0 to 5i%. Grade of main gangway, tV to to of i%, 
in favor of load. Trains of 10 cars are hauled on main 
gangway, and 4 cars on the slope; weights of cars same as 
above. 

Locomotive-tank pressure at start, 600 lbs.; at end of trip, 
200 lbs. Average working pressure, 180 lbs. The cost of the 
plant was as follows : 

One Norwalk 3-stage compressor, erected $5,180. 74 

Pipe-line, 4,200 ft., 5 in., including 3 charging stations 2.951 .06 

Two Baldwin compressed-air locomotives and fittings 4,904.33 

Alterations in gangways to adapt them to locomotive haulage 665.17 

Total cost $13,701 . 30 

Daily operating cost, for 180 days in the year $14. 69 

Fixed charges, depreciation, repairs, etc., figured at 10 per cent, 

together with cost of steam power 9 . 00 

Total running expenses per day $23 . 69 

Haulage cost per car, at 660 cars per day 3.6 cents 

Previous cost of mule haulage per car 5.1 " 

Saving per year, about $1,800. 00 

5. At the Wilson Colliery, of the D. & H. Coal Co., a large 
locomotive was installed by the Dickson Manufacturing Co., 
having six 26-in. drivers; wheel-base, 7 ft.; cylinders, 9 by 14 
ins.; gage of track, 30 ins. The locomotive carries two tanks, 
18 ft. 6 ins. and 15 ft. 6 ins. by 30 ins. diameter; capacity, 160 
cu.ft. of air at 600 lbs. Pipe-hne, 4,100 ft. long; pressure, 700 
lbs. Total charging time, i min. 25 seconds. After reduction to 
125 lbs. working pressure, the air is reheated. Trains are usually 
of 30 cars, each weighing loaded, 5,850 lbs. (the locomotive has 
a capacity of 50 cars). Grades from J of 1% against the load, to 
1% in favor. Round- trip time, for 8,200 ft. plus a switching 
distance of 800 ft., 16 min. Cost of haulage per ton-mile, gross, 
about ij cents. 

6. The Anaconda Copper Mine, Butte, Mont., has a number 
of compressed-air locomotives with 5 by lo-in. cylinders and 



462 COMPRESSED AIR PLANT 

weighing 10,000 lbs. Over-all dimensions: height, 58 ins.; 
width, 58 ins.; length, 10 ft. 4^ ins.; 4 driving wheels, 23 ins. 
diameter; wheel-base, 36 ins., designed for curves of 12-ft. radius. 
Capacity of main tank, 47 cu.ft.; pressure, 550 lbs.; working 
pressure, 125 lbs.; charging time, 60 seconds. Length of haul, 
2,400 ft. round trip; load, 6 cars, weighing loaded 3,450 lbs. 
each; track nearly level. The locomotives can make 2 round 
trips, or 4,800 ft. on i charge, with cold air; by reheating with 
hot water, 3 round trips can be m^ade. 

At the Anaconda reduction works, there are 13 H. K. Porter 
locomotives, employed in handling the products between the 
different divisions of the plant; length of haul, 1,000-7,000 ft. 
Twelve locomotives have the following dimensions; weight, 
26,000 lbs.; cylinders, 9I by 14 ins.; driving wheels, 28 ins.; 
wheel-base, 54 ins.; main tanks, 132 cu.ft.; drawbar pull, 
5,700 lbs. Another locomotive weighs 42,000 lbs.; cylinders, 
12 by 18 ins.; driving wheels, 36 ins.; wheel-base, 60 ins.; main 
tanks, 218 cu.ft.; drawbar pull, 9,180 lbs. Tank pressure, 
700-800 lbs.; working pressure, 150 lbs. 

7. The Homestake Mining Co., Lead, S. D., employs under- 
ground 10 H. K. Porter locomotives, weighing 9,500 lbs. and 
m-casuring over all, 4 ft. 11 ins. high by 3 ft. 3I ins. wide by 10 ft. 
6 ins. long. Gage of track, 18 ins. They have a detachable 
rear end (as in Fig. 234) to permit of transferring them from 
level to level, on a cage with a 9-f t. platform. 

8. At the Aragon Iron Mine, Norway, Mich., is an H. K. 
Porter locomotive. Weight, 7 tons; height, 5 ft. 2 ins. ; width, 
4 ft. 2 ins. ; length, 12 ft. over all. Four 24-in. drivers; wheel- 
base, 48 ins.; gage, 22^ ins.; cylinders, 7 by 12 ins.; tank pres- 
sure, 700 lbs. ; working pressure, 140 lbs. ; charging time, 30-60 
seconds. Haulage distance, 1,200-4,000 ft. ; pipe-line, 1,800 ft. 
Locomotive hauls four 20-car trains per 10 hours, from each of 
10 loading places. Weight of loaded train, including locomotive, 
43 tons; weight empty train, 18 tons. At the compressor are 
2 receivers, each 3 by 17 ft. 

9. Compressed-air haulage plant at No. 6 Colliery, Susque- 
hanna Coal Co., Glen Lyon, Penn. Following is an abstract 



COMPRESSED AIR HAULAGE 463 

of tests by J. H. Bowden and R. V. Norris.* Though the plant 
is old, the results are useful. 

Compressor: Norwalk, three-stage; steam cylinder, 20 by 24 
ins.; air cylinders, 12 J, 9-2 and 5 ins. by 24 ins.; capacity, at 100 
revolutions, 296 cu.ft. free air per min., compressed to 600 lbs. 
Main pipe-line at No. 6 shaft, 4,380 ft. long, 5 ins. diameter, with 
5 charging stations, and capacity of 608 cu.ft. Branch line, in 
No. 6 slope, 3,100 ft. long, 3 ins. diameter, with 3 charging sta- 
tions, and capacity of 159 cu.ft. 

Locomotives: two, by H. K. Porter Co.; weight, 8 tons; tank 
capacity, 130 cu.ft.; pressure, 550 lbs., reduced to 160 lbs. in an 
8-in. auxiliary reservoir of 4.2 cu.ft. capacity. Cylinders, 
7 by 14 ins. ; four 24-in. drivers. 

At No. 6 shaft the run averages 4,000 ft. each way, on J-2f% 
grades, averaging about 1%, in favor of the load. Run at No. 6 
slope averages 2,100 ft., with nearly the same grades. Cars 
weigh 2,800 lbs. empty, and about 9,800 lbs. loaded; trains, 
12-20 cars. The shaft locomotive hauls about 355, and the 
slope locomotive 320 cars, per 10 hours, doing the work of 32 
mules. Tests on the compressor showed 150 indicated H.P. at 
131 revs., compressing 387.8 cu.ft. free air per min. 

The combined capacity of both pipe-lines is 767 cu.ft., 
which, at 600 lbs. gage pressure, is equivalent to 32,500 cu.ft. 
free air. Capacity of locomotive main and auxiliary tanks, 
134.6 cu.ft.; at 508 lbs. (at which pressure the tanks equalize 
with the mains, the initial pressure being 600 lbs.) , this is equiv- 
alent to 4,845 cu.ft. free air. In standing 12 hours, the pipe- 
line pressure falls to 350 lbs.; hence the loss per hour from leak- 
age is 974 cu.ft. free air, or 4.18% of total air compressed. 

Average volume free air used by both locomotives per ton- 
mile: gross, 100 cu.ft.; net, 180 cu.ft. Another test showed a 
total consumption of 223,020 cu.ft. free air, for hauling 676 cars 
per day. The volume of free air apparently compressed for this 
work was 279,200 cu.ft., of which 83.4% is accounted for, leaving 
16.6% for leakage and slip in the compressor, leakage in air 
lines, and changes in temperature. 

* Transactions American Institute of Mining Engineers, Vol. XXX, p. 566, 



464 



COMPRESSED AIR PLANT 



Table LIII 
Haulage Data, Glen Lyon Colliery 



Number of trips, empty 

Number of trips, loaded 

Average number cars per trip, empty . . 

Average number of cars per trip, loaded 

Average cu.ft. free air per trip, empty 

Average cu.ft. free air per trip, loaded 

Average cu.ft. free air per round trip 

Average cu.ft. free air per ton-mile, on gross tonnage. 
Average cu. ft. free air per ton-mile, on net tonnage. 



Sh.aft Loco. 



113 

203 



No. 2 


No. 3 


01 ope 
Loco. 


Plane. 


Plane. 




3 


10 


16 


3 


10 


15 


15-33 


12.7 


II. 4 


13 


13 


II 3 


1,724 


5,686 


1,230 


1,631 


1,898 


599 


3,355 


7,584 


1,829 



71 
128 



Cost of the plant, omitting boilers, was: 

Compressor and extras $2,955 . 75 

Two locomotives and extras 5,869. 76 

Pipe-line: 5-in. line, 6,000 ft $2,914.32 

3-in. line, 4,000 ft 1,240.46 

4,154 78 

Steam connections to compressor 278. 27 

Material and labor on compressor house and foundations, and install- 
ing pipe-line, etc , 1,525 . 23 

Charging stations 372.21 

Total cost $15,156.00 

Average cost of operation for 2 years, on basis of 170 days* 
work per year, was $12.60 per lo-hour shift, including $2.32 for 
steam for compressor, furnished by main boiler plant of mine. 
Adding proportion of fixed charges, with interest, depreciation 
and repairs, the daily cost (300 days' work per year) would be 
$18.52 per day. For the 2 years, the average cost per ton-mile 
was as shown in Table LIV. 

In these two years the saving over the expense of the mule 
haulage, previously employed, was $14,218, or nearly the total 
cost of the haulage plant. 



COMPRESSED AIR HAULAGE 



465 



Table LIV 
Haulage Operating Costs, Glen Lyon Colliery 







1897 (179 Days). 


1898 (160 Days). 




Daily 
Ton- 
Miles. 


Daily 
Cost. 


Cost per 
Ton- 
Mile, 
Cents. 


Daily 
Ton- 
Miles. 


Daily 

Cost. 


Qost per 
Ton- 
Mile, 
Cents. 


Shaft locomotive, 


gross tonnage. 


1,485 


$11.12 


0.75 


1,521 


$12 .00 


0.79 


Shaft locomotive, 


net tonnage. . 


825 


II . 12 


1-35 


845 


12 ,00 


1.42 


Slope locomotive, 


gross tonnage. 


648 


II . 12 


1.72 


720 


12 .00 


1.67 


Slope locomotive, 


net tonnage . . 


360 


II . 12 


3 09 


400 


12.00 


3.00 


Both locomotives 


gross tonnage 


2,133 


22.23 


1.05 


2,241 


24.01 


1.07 


Both locomotives 


, net tonnage. 


1,185 


22.23 


1.89 


1,245 


24.01 


1-93 



10. Following is the cost of a large colliery p'ant, as given by 
Beverly S. Randolph,* who installed and operated it: 

Three-stage, compound compressor $5,300 

Pipe line: 5,600 ft., 5-in $5,600 

3,100 ft., 3.V-in V 1,700 

1,000 ft., i^-in 300 

7,600 

Two main locomotives, weight 30,000 lbs. each 6,000 

Five gathering-locomotives, weight 8,000 lbs. each 10,000 

Two boilers, each 80 H.P i ,000 

Installation, labor, and material 4,000 

Total cost $33,900 

This plant includes an unusually large number of small 
gathering-locomotives. If the equipment had consisted of four 
25,000-lb. engines, costing, say, $2,800 each, and which would 
do the same work, the total cost would be $29,100. 

J I. Inspiration Consolidated Copper Co., Arizona. In 1914 
six 10- ton two-stage locomotives were installed for under- 
ground service; in August, 191 7, there were 11 locomotives in 
use, including 2 spares. Over-all dimensions, 5 ft. 9 ins. high by 
5 ft. 4 ins. wide by 16 ft. long. Rigid wheel-base, 4 ft.; track 
gage, 30 ins.; 4 drivers, 26 in. diameter; total weight, 20,000 

* Trans. Instn. of Min. Engrs. (England), Vol. XXVII (1904), p. 433. 



466 COMPRESSED AIR PLANT 

lbs., all •on drivers; drawbar pull, 4,400 lbs.; cylinders, 7 ins. 
and 14 ins. by 14-in. stroke; locomotive storage tank, 40-in. 
diameter by 12 ft. 8 ins. long, 105 cu.ft. capacity. 

Air at 1,000 lbs. is furnished by two 4-stage compressors, 
each having a capacity of 1,125 cu.ft. of free air per min. Main 
air-line in shaft is 6 in., with 3 -in. and 2-in. branches to charging 
stations; there are no receivers. Average charging time, i min. 
Charging pressure, 800 lbs., which is dropped by a reducing 
valve to 250 lbs. in an auxiliary reservoir. In the high-pressure 
cylinder, the air is expanded to 50 lbs., causing a temperature 
drop to about 140° F. below normal. Before the air enters the 
low-pressure cylinder, its normal temperature is approximately 
restored by an interheater. 

Trains average 15 cars (maximum, '25 cars), each weighing 
2 tons and carrying 5 tons of ore; train crew, engineer and 2 
helpers; 5 cars at a time are dumped in a rotary tipple. Average 
haul (August, 191 7), 0.475 niile. Locomotive is charged once 
per round trip. Average number of trips, 15 per 7 hours of 
actual hauling time. Power per ton-mile, 0.796 kw. Cost of 
haulage is variable, due to sliding wage scale. 



CHAPTER XXVII 

MEASUREMENT OF AIR CONSUMPTION 

A DISCUSSION of the transmission of compressed air through 
long pipes, with formulas and numerical data, is given in Chapter 
XVI. The behavior of air is different when flowing through short 
pipes, or orifices in thin plates, and under pressures relatively 
small as compared to those existing in ordinary compressed-air 
transmission (see page 222). 

Flow of Air through Orifices or Short Tubes. The funda- 
mental equation isv = c\ 2gh, in which: v is the velocity of flow, 
feet per second; g, the force of gravity, or 32.2; h, the height in 
feet of a column of air, or head, required to produce the pressure 
under which the air is flowing, that is, h = p^ — p^ = diEerence 
between the pressures on the two sides of the plate containing 
the orifice; and c, a coefficient of flow, determined experiment- 
ally. The volume discharged, in cubic feet per second, is equal 
to V multipHed by the area of the orifice in square feet. 

Values of constant " c '^ 

(Weisbach, Mechanics of Engineering, p. 945-947). 

FLOW THROUGH AN ORIFICE IN A THIN PLATE 
Diameter of orifice =1 cm. = 0.394 inch: 

Ratio of pressures on the two sides 

of the orifice = — 1.05 1.09 1.43 1.65 1.89 2.15 

Values of c o-555 o-S^Q 0.692 0.724 0.754 0.788 

Diameter of orifice =2.14 cm. = 0.843 inch: 

Ratio of pressures 1.05 1.09 1.36. 1.67 2.01 

Valuesofc 0.558 0.573 0.634 0.678 0.723 

467 



468 COMPRESSED AIR PLANT 



FLOW THROUGH A SHORT TUBE 

Diameter of tube=i cm. = 0.394 inch; length 3 cm. = 1.181 inch: 

Ratio of pressures = -^ 1.05 i.io 1.30 

Values of f: 0730 0.771 0.S30 

Diameter of tube=i.4i4 cm. ^0.557 in.; length, 4.242 cm. = 1.670 inch: 

Ratio of pressures i . 41 i . 69 

Values oi c 0.813 0.822 

Diametef of tube=i cm. 0.394 in.; length, i .6 cm. = 0.630 in. Edge of orifice 

rounded : 

Ratio of pressures 1.24 1.38 1.59 1.85 2.14 

Values of c 0.979 0.986 0.965 0.971 0.978 

In this case c always approximates unity. A short conical pipe, rounded 
off at the inlet orifice, gave nearly the same values for c. 

An empirical formula for the velocity of flow, suitable for 
dealing with the small differences of pressure used in ventilation, 
is (Clark's Rules, Tables and Data. p. 891) : 

F = cV?«^X773-x(x+'-^^)x^ 
12 " V 493 / P 

= 352C\[i +0.00203(^-32)]- 

where: F = velocity of flow, feet per second; C = coefficient of 
efflux; ^ = 32.2; /? = inches of water column measuring the dif- 
ference of pressure; / = the temperature, Fahr.; /? = barometric 
pressure, in inches of mercury. The constant 773.2 is the vol- 
ume of air at 32°, under a pressure of 29.92 inches of mercury, 
vhen the volume of an equal weight of water is taken as i. 

For 62° F., F = 363C\-, andif /? = 29.92, V = ()6.y^Cy/lt, 

In this formula, the values of C (Weisbach), for pressures of 
0.23 to I.I atmosphere, are: 

For circular orifices in thin plates. o. 56 to o. 79 

For short cvhndrical tubes 0.81 to 0.S4 



MEASUREMENT OF AIR CONSUMPTION 



469 



For the same, rounded at the inner end 0.92 to o. 93 

For conoidal mouthpieces, in form of the "vena 

contracta " o . 97 to o . 99 

For conical converging mouthpieces o . 90 to o . 99 



Table LV 

Actual Discharge, Pounds of Dry Air per Second, at 60"^ F. 
AND 14.7 Lbs. Barometric Pressure, for Circular Ori- 
fices IN A Plate 0.057 In. Thick (i Lb. of Air under 
the above Conditions = 13.09 Cu. Ft.).* Kent. 



Pressure, 






Diameter of 


Orifice 


Inches. 








Inches of 

Water 
Column. t 


•3125 
001 14 


. 500 
0.0293 


I . 000 ] 


[.500 


2.000 


2,500 3 


.000 3 


.500 4 


.000 


1 
2 


.0117 


0263 


.0468 


•073 


105 


143 


187 


I 


00162 


.00416 


.0166 


0373 


.0663 


.103 


149 


202 


264 


l| 


00199 


.00510 


.0203 


0457 


.081T 


.127 


182 


248 


323 


2 


00231 


.00590 


•0235 


053S 


•0937 


. 146 


210 


285 


373 


^1 
23 


00259 


.00662 


.0263 


0591 


.1050 


.163 


235 


319 


416 


3 


00285 


.00726 


.0289 


0648 


.1150 


.179 


257 


349 


455 


^1 
d2 


00308 


.007S6 


.0312 


0700 


. 1240 


•193 


277 


377 


491 


4 


00330 


.00S42 


•0334 


0749 


■1330 


. 206 


296 


402 


525 


4l 


00351 


.00895 


•0355 


0794 


. 1410 


. 219 


314 


426 


556 


5 


00371 


• 00945 


•0375 


0838 


. 1480 


•231 


331 


449 


586 


S'i 


00390 


•00993 


•0393 


0879 


•1550 


. 242 


347 


471 


613 


6 


00408 


.01049 


.0411 


0918 


. 1620 


.252 


362 


492 


640 



* The general expression for the weight of dry air is: W= ~~^ — . where W— weight of 

I cu.ft. of air; B = absolute pressure, inches of mercury; T= absolute temperature (F.). 
t I inch of water column =0.0361 lb. per sq.ixi. =5.20 lbs. per sq.ft. 

Measurement of Low-pressure Air. Fig. 247 shows an appa- 
ratus for measuring small quantities of air at low pressure.* 
It can be used for any ordinary working pressure, as employed 
for mine service, but the actual measurement is applied to an 
equal flow of air under a low pressure. That is, for convenience 
in measuring while carrying on the test, an equal volume of flow 
of low-pressure air is substituted for the regular high-pressure 

* G. S. Weymouth and C. C. Freeman. Jour, Chamber of Mines, Kaigoorlie, 
Western i\.ustralia, 191 2. 



470 COMPRESSED AIR PLANT 

supply. The apparatus here described is for dealing with 
small volumes, say 20 cu. ft., or less, of free air per minute; 
but, by using larger tanks, any quantity can be measured, pro- 
vided the supply pressure is constant during the test and the 
flow regular for the few seconds necessary to obtain the first 
gage reading. The volume of air thus isolated is then measured 
at leisure. In using this method, the regular supply must be 
diverted for about 0.5 minute. 

Connected with the air supply is a vertical i-in. pipe, with 
globe-valves A and D, stop-cock B, and a mercury gage /. This 
pipe leads by a hose to a closed water tank K, of about 30 cu. ft. 
capacity, which is provided with a gage glass and a short open 
hose fl" from the bottom. A branch pipe, with stop-cock C and 
globe-valve E, leads to a 5-gal. drum (of light iron or tin plate) 
connected with which is a water-gage /. The drum has an 
opening at F, to which disks with orifices of different sizes can 
be applied, the area of orifice being such as to keep the pressure 
in the drum low enough to eliminate factors caused by varia- 
tions of volume due to pressure. These orifices may be of, say, 
J, J, h and I sq. in area. 

To make a test, valve A is opened to admit the necessary 
quantity of air; cock C is then opened and valve E regulated to 
give any required pressure on the mercury gage / (for example, 
5 lbs. per square inch). An orifice is attached at F to the 5-gal. 
drum, to give a reading on the water-gage / of 2 to 9 ins. 
( = 0.0722 to 0.325 lb. per square inch). The hose at G is dis- 
connected, cock B opened, cock C closed, and valve D regulated 
to give the same reading on gage / as before ; thus, the resistance 
of valve D is made equal to that of valve E, leading to the orifice 
at F. Tank K is then nearly filled with water, and its level noted 
on the gage glass; cock C is opened, B is closed, and the hose is 
connected at G to tank K. Next, cock B is opened and C closed 
simultaneously, and the free end of hose H lowered ; the reading 
on gage / is noted and, when the tank is nearly empty, C is 
opened and B closed simultaneously. The time the air is passing 
to tank K is taken by stop-watch; the air in K is reduced to 
atmospheric pressure (by bringing the free end of hose H to the 



MEASUREMENT OF AIR CONSUMPTION 



471 



water level in the tank), and the volume of air noted. If neces- 
sary, due to any back pressure from tank K, gage / is brought 
to the same level as when air was passing to K, by manipulating 
valve E, and a reading on / is taken. 

Table LVI shows the results of measurements, using orifices 
of 0.214, 0.441 and 0.797 sq. in., a pressure on the mercury gage 



Hose n ^- e 
^— -— — —-v Opening for 

U ^ Orifices 

^^ E 
I'Valve 



lis/ 

J 
Water 

Oage 




V4 Hose 



f-Hose Clamp 



Mercury 
Gage 



K 

Closed Tank, 

about 30 cu, ft. 

Capacity 



.5^ 




1 Valve 



From Air Main, 
80 lbs. Pressure 



Fig. 247.- 



-Diagram of Air Apparatus for Measuring Low-pressure Air (Weymouth 
and Freeman). 



equal to 9 lbs. per square inch, and volumes of free air of 9 to 16 
cu. ft. per minute. The volumes discharged through the orifices, 
computed from the formula v = c\/ 2gh, check closely with those 
measured in the tank. In using the formula, the coefficient of 
discharge was taken as 0.64 and the weight of i cu. ft. of air at 
62° F., as 0.0671 lb. 

Measurement of High-pressure Air. (i) For comparatively 
small volumes, the air deUvered by a compressor may be passed 



472 



COMPRESSED AIR PLANT 



into a tank or receiver of known capacity, provided w^ith a pres- 
sure gage; (2) the air may be measured by causing it to displace 
water in a tank; (3) the air may be metered. 

Table LVI , 

Results of Measurements of Low-pressure Air by Appa- 
ratus OF G. S. Weymouth and C. C. Freeman. {Mining 
Press, April 20, 191 2). 



Area of 


Mercury 

Gage /, 

Ins. 


Water 

Gage J , 

Ins. 


Calculated Vol- 
ume. COEFF. OF 
DlSCH.= 0.64- 


Volume 

Passed 

into 

Tank. 

Cu. Ft. 


Time, 
Sees. 


Volumes of 

Free Air in 

Tank. 


Orifice F, 




Sq. In. 


Cu. Fr. 
per Sec. 


Cu. Ft. 
per Min. 


Cu. Ft. 
per Sec. 


Cu. Ft. 
per Min. 


0. 214 


12 


q 9 
"16 


O.1S4 


II .0 


17 -49 


96.5 


O.181 


10.9 


0. 214 


T9 


9^ 


0.194 


II. 6 


22> 


01 


108.0 


0.212 


12. 7 


0. 214 


16 


„ 1 
/ 16 


0. 167 


10. 


16 


II 


94-5 


O.171 


10.3 


0. 214 


16 


58 


152 


9.1 


13 


82 


88.5 


0.156 


9.4 


0. 214 


16 


_3 
4 


O.151 


9.1 


13 


82 


88.0 


0.157 ! 9.4 


0.441 


9 


2\ 


0.194 


II. 6 


19 


3>3> 


96.0 


0. 201 


12. I 


0.441 


12 


2>\ 


0.237 


14.2 


20 


48 


86.0 


0.238 


14 3 


0.441 


13 


IH 


0.186 


II . 2 


22 


55 


116. 


0.194 


II. 6 


0.441 


16 


4i 


0. 271 


16.3 


22 


86 


87.0 


0.263 


15.8 


0.441 


16 


a\ 


0. 267 


16.0 


19 


40 


74 


0. 262 


15 7 


0.797 


12 


15 
16 


0.228 


13-7 


20 


-iS 


86.0 


0.238 


14 3 


Aver'g'?. 






0.203 


12. 2 






0. 207 


12.4 





Measurement of Discharge through Orifices. This is as 
accurate as the use of tanks, and is applicable to both small and 
large volumes. The compressed air is discharged through orifices 
of known area, and the equivalent volume of free air computed. 

One apparatus for this method consists of a short piece of 
wrought-iron pipe, in which are set a number of short branches, 
each pro\'ided with an orifice, the orifices being of different diam- 
eters.* Each branch has a valve, the whole resembling a mani- 
fold as used for connecting a group of rock drills to an air main. 
Fig. 248 shows a meter with 8 orifices, ranging from ^ to f in. 



* F. D. Holdsworth, Chief Air-Compressor Engineer, Claremont plant, Sulli- 
van Machinery Co. Eng. and Min. Jour., May 25, 191 2, p. 1028. 



MEASUREMENT OF AIR CONSUMPTION 



473 



diameter. The orifices are accurately reamed holes in disks of 
steel plate, held by flanges in the outer ends of the branch pipes. 
For the larger orifices the disks are J in. thick; for the smaller, 
f in. The back or pressure side of each hole is rounded to a 
radius jt in. less than the thickness of the plate. The actual 
diameter of the finished hole is measured by a micrometer and 
the area calculated. 




Fig. 248. — Manifold and Nozzles for Measuring the Discharge of Compressed Air 

{En g. and Mm. Jour., May 25, 1912). 

The rate of flow through an orifice of this form is obtained 
from Fliegner's formula, 

G = o.53— =- 



where G = flow, pounds per second; ^=area of orifice, square 
inches; P and r = absolute pressure and temperature (F.) of 



474 COMPRESSED AIR PLANT 

the air behind the orifice. For the formula for the weight of a 
cubic foot of air, see the footnote under Table LV. 

Several small orifices are used instead of a single large one, 
because, though the formula is accurate for orifices not greater 
than I in., it is inaccurate for larger ones. Also, by ha\nng a 
series of orifices of different diameters, any desired combination 
may be made according to the volume of air to be measured. 
Tapped into the main pipe of the manifold is an accurate pressure 
gage and a well for a thermometer, w^hich, for a 2-stage com- 
pressor, should read to at least 300° F. 

To prepare for a test, disconnect the air main near the receiver 
and attach the manifold. Stop all other outlets from the receiver, 
so that all the air entering the receiver will pass through the 
orifices. Place another thermometer in the compressor intake, 
close to the cyHnder. The compressor speed is recorded by a 
revolution counter. Determine by trial the proper combination 
of orifices to maintain the desired pressure. Run the com- 
pressor about 2 hours, discharging through the orifices, so that 
both pressure and temperature will become maximum and con- 
stant. 

One observer begins recording the compressor revolutions, 
and another, at i-minute intervals, takes the temperature and 
pressure readings at the orifices; at the same time, the intake 
temperature is also recorded. In 10 or 15 minutes enough tem- 
perature and pressure readings will be obtained, and a final 
reading is made of the revolution counter, to find the total 
number of revolutions for the period of the test. The total 
piston displacement is then computed. Finally, the quantity of 
air discharged through the orifices, determined by the formula 
and divided by the total displacement, gives the volumetric 
efficiency. 

For accurate results, obtain the local barometric reading, 
from the nearest Weather Bureau station, or otherwise. This, 
with the temperature at the compressor intake, is used in the 

1,72 kB 

formula, W= ^^— (see note under Table LV), for reducing 

to cubic feet of free air per minute the pounds of air per second, 



MEASUREMENT OF AIR CONSUMPTION 475 

computed by Fliegner's formula. In testing a stage compressor, 
only the displacement of the low-pressure cylinder is used, since 
this measures the volumetric capacity of the compressor. 

A method of measuring air by pitot tubes was adopted at a 
mine in northern Michigan, for finding the quantity of air used 
in a large blacksmith shop.* The compressor was working at a 
pressure of 75 to 80 lbs. 

Two pitot tubes, made of copper tubing re in. outside diam- 
eter, were soldered into a f-in. to |-in. bushing. This bushing 
in turn was screwed into a hole tapped in the side of a piece of 
i-in. iron pipe, 20 ins. long, which was inserted in the air main. 
The pitot tubes projected J in. into the pipe; one of them being 
directed toward the air current, the other set at right angles to it. 
They were connected to a U-tube, made of two gage glasses 
joined at their lower ends by rubber hose. Either a water or a 
mercury gage may be used to determine the velocity head of 
the air current. The 20-in. piece of i-in. pipe was provided with 
a thermometer and pressure gage, set behind the pitot tubes and 
at a little distance from them, to avoid producing eddies. 

The volume of air flowing through the i-in. pipe was computed 
by the formula 

where (3 = cubic feet of free air per minute, at 60° F. and 28.5 ins. 
barometer ; H = height of water column, inches, representing the 
velocity head (if mercury be used, multiply the reading by 13.6, 
which is the specific gravity of mercury) ; P = reading of pressure 
gage; / = thermometric temperature of the compressed air 
(degrees F.). 

To furnish some storage capacity, and so prevent sharp 
peaks in the flow of air caused by the intermittent use of the drill 
sharpener and the shop hammer, the measuring device was set 
in the air fine at a Httle distance from the shop. Measurements 
were made on a Saturday afternoon, when air was being used 
only for the shop and rock-houses. The result showed: total 

* B. B. Hood, Eng. and Min. Jour., June 27, 1014, p. 1283. 



476 co:mpressed air plant 

free air compressed, 670,000 cu. ft.; blow-off, 338,000 cu. ft.; 
line loss, 118,000 cu. ft.; blacksmith shop, 27,000 cu. ft.; rock- 
houses, 187,000 cu. ft. The high Hne loss was largely due to 
expansion and contraction in the spiral-riveted pipe and its 
joints. Most of the leakage was in the 500 ft. of air main 
nearest the compressor, where the variations of temperature 
were naturally greatest. Underground, the piping was found to 
be practically tight. 

* Ready-made " Meters of several forms are on the market; 
for example, the ''Tool-om-eter,'^ made by the New Jersey 
]\Ieter Co., Plainfield, N. J., and the instrument of the Bailey 
Meter Co., Boston. 

The Bailey meter works on the principle of measuring accu- 
rately the pressure difference across an orifice placed between 
a pair of flanges in the pipe line. A special orifice plate is used, 
made of y^^-in. Monell metal. The area of orifice is proportioned 
to the size of pipe, so that, at average rates of flow, a pressure- 
drop of about J lb. takes place. The record is automatically 
made on a direct-reading chart, with uniform graduations. 
(For a detailed description, see Met. b' Chem. Eng'g., April 15, 
1916, p. 456.) 

The " Tool-om-eter " (Fig. 249) is a small, simple direct- vol- 
ume gage, employing the principle of multiple nozzles. It has 
but one moving part, which consists of the weighted piston a, in 
the metering cylinder h, and a piston rod c, which carries a small 
piston d in the oil dashpot e. An extension of the piston rod 
above a moves freely, without contact, inside of the sight glass /, 
and the height of its upper end is read on the accompanying scale, 
which records cubic feet of free air per minute. 

At g air enters the space around the dashpot cylinder e and 
the oil reservoir, and then passes through ported openings into 
the metering cylinder h. Through the walls of this cyhnder are 
drilled a large number of small holes, accurately spaced and 
reamed. To pass to the outlet chamber h, the compressed air 
lifts piston a, thus exposing some of the holes to the flow. The 
air leaves the meter at i. 

The small ''head" or difference of pressure (a few ounces per 



MEASUREMENT OF AIR CONSUMPTION 



477 



squ8.re inch) , produced between the interior of cylinder b and the 
chamber h, is fixed by the relation of the weight of the moving 
element to the area of the piston a on which the difference of 
pressure acts. That is, the moving 
element rises until its weight is exactly 
supported by the differencepf pressure; 
so that the pistons and rod float in 
static balance, in aposition correspond- 
ing to the volume of air flowing and 
the number of "holes exposedincyHnder 
b. The scale is calibrated by com- 
parison with a standard instrument. 

This meter is made in two sizes: 
the smaller for flows of lo to loo cu. 
ft., and the larger for 50 to 300 cu. ft. 
of free air per minute. 

It may be suggested that the 
water meter m.ade by the Worthing- 
ton Pump & Machinery Co. might 
readily be modified for measuring 
compressed air. The measuring cham- 
ber of this meter contains a ''wabbling 
disk," resting at its apex on a ball, 
and set between the inlet and dis- 
charge openings. A slit cut in one 
side of the disk fits a radial septum or 
division in the disk chamber; water 
is admitted on one side of the septum 
and discharged at the other side. 
The water pressure causes the disk to ^^ 
make a uniform wabbling motion, t. ^, u^t^ 1 

° Fig. 249. — The "Toolom-eter, 

which passes water continuously for Measuring Compressed Air. 

through the meter; though there is 

at no time any free opening, through which water could flow 
without moving the disk. A short rod, attached at right angles 
to the upper side of the disk, com.municates motion to a gear- 
train, which operates a counter registering cubic feet of water on 




478 COMPRESSED AIR PLAN'T 

a dial. This meter, like some others, is of the ''inferential " t}pe; 
that is, it must be rated to determine its discharge. 

Measuring Tanks. Two large vertical tanks are generally 
used, connected with the compressor and with a pump for 
supphdng water to the tanks, long vertical gage glasses being 
provided for reading the water level. The air in the two tanks 
is always separated by a piston-like mass of water, which flows 
back and forth from one tank to the other. The change in level, 
multiplied by the area of the tank, is equal to the volume of air 
that has flowed out. In order that the observed differences of 
water level shall accurately represent corresponding volumes of 
air, it is evident that the internal cross-sections of the tanks 
must be constant throughout their heights. 

The machine to be tested is connected to one tank, both 
tanks are filled with air at the desired pressure, and the pump 
keeps the air pressure constant by forcing in water. On starting 
the drill, or other machine, the height of water in the gage glass 
is noted. After a run of sufficient length to secure accuracy, the 
drill is stopped and the height of water in the gage glass on the 
tank which has been used is again noted. The increase in volume 
of water in the tank is the volume of compressed air used during 
the run; from this the volume of free air is computed. Having 
two tanks, one is used while the other is being filled with air, 
thus allowing tests to be run for any desired time. 

Fig. 250 shows an apparatus, which though more elaborate is 
practical and convenient, and the tanks need not be of uniform 
cross-section throughout.* Instead of using long gage glasses, 
reaching from top to bottom, one tank has two short glasses, / 
and h, one at top, the other at the bottom; the companion tank 
has a single gage, at the top. 

As the tank air must be kept at constant pressure, a reducing 
valve is provided (the valve known as the "5-4 locomotive 
governor," made by the Westinghouse Air Brake Co., is 
recommended). It is arranged as at a. Fig. 250. The control 
pipe h should not be connected directly to either tank, as 
the valve continuously exhausts air to the atmosphere. A 

* Walter S. Weeks, Mining Press, December 15, 191 7. p. 855. 



MEASUREMENT OF AIR CONSUIMPTION 



479 



lubricator r is placed on pipe h, for feeding a thin oil to the 
valve. 

To transfer the air from one tank to the other, a 4-way valve 



o 




Blank Plftnge 






Q 



StoB^^ 



"Water Falling 



=&: 




ti 



"iinPRto UrUl_ 



^M) 





II 



-Slop 



Water-Hisinff 



Fig. 250. — Measuring Tanks for Rock-drill Testing (W. S. Weeks, Mining Press, 

December 15, 1917). 



is generally used (see Fig. 251). It has the disadvantage of being 
liable to leakage. Fig. 250 shows an arrangement of 4 pipes, 
each provided with a single, i-in. cut-out valve c. The valve 
levers are connected, so that all are thrown together, thus pror 



480 COMPRESSED AIR PLANT 

ducing the effect of a 4-way valve. In pipe d, connecting the 
tanks at the bottom, is a quick-closing gate- valve e. 

The volume of one tank only needs to be measured, that 
having the two gage glasses. First place a mark on glass /, 
and fill the tank with water to that point through pipe g. Close 
valve e, and draw off water through pipe g until the water level 
is sighted on glass h. Then carefully draw off more water, 
until a convenient even number of cubic feet is reached, and 
mark the corresponding point on glass //. Thus, when in opera- 
tion the water rises from the lower to the higher mark, the 
measured volume of air has been displaced. 

To test a drill, connect the hose at /, as shown ; then fill tank 

1 with water; close the gate- valve e\ bring the water level in 
tank II to the lower mark ; open the gate-valve and throw valves 
c to the position shown, thus forcing water from tank I to tank II 
until it reaches the upper mark; then close the gate- valve. The 
reducing valve a is adjusted to hold the pressure constant, not- 
withstanding the small changes in hydraulic head on the air. 
When necessary, the air in the tanks may be exhausted to the 
atmosphere through the 3-way valve m. The drill test may 
be continued as long as desired by running the water a number 
of times from one tank to the other. For ordinary testing, 
the volume between the gage -glass marks on tank II should be 
16 or 18 cu. ft. 

A less expensive apparatus consists of two small tanks, say 

2 ft. diameter by 4 or 5 ft. high, connected at the bottom by a 
3-in. pipe and provided with gage- glasses (Fig. 251).* The 
tanks are about half filled with water. Pipes from the top of 
each tank are connected to a common 4-way valve. With the 
valve in the position shown in the cut, water in the right-hand 
tank is forced by the air pressure into the left-hand tank, and 
the air in the latter is driven out through the valve to the drill. 
When the water level reaches a definite point near the upper end 
of the gage-glass on the left-hand tank, the valve handle is 
quickly thrown to the position shown by dotted lines. The 

* George IT. Oilman, Sullivan Machinery Co., Compressed Air Magazine, 
Aug. 1912, p. 6510. 



MEASUREMENT OF AIR CONSUMPTION 



481 



process is thus reversed, the air in the right-hand tank then 
passing to the drill. 



1 - Way Valve 




Fig. 251. — Tanks for Measuring Air Consumption of Rock Drills. 

The total volume of compressed air used is found by counting 
the number of times each tank is filled and emptied during a 
given time, and the corresponding volume of free air is com- 
puted as already explained. 

Measuring Air Discharge with One Tank.* An accurate 
method is as follows. To one end of the air receiver is connected 



Pressure 



Well for 
Thermometer \ 

Valve Valve 
Discharge to A^_^ 

Atmosphere ■^^c^-O^^—i^'- 




From 



Compoessor 



Fig. 252. — Air Receiver Arranged for Measuring Air from a Compressor. 

the air pipe from the compressor. At the other end is a hori- 
zontal discharge pipe, provided with two globe valves set close 
together and a well for a thermometer (Fig. 252). This pipe 

* E. E. Fessenden, Compressed Air Magazine, February, 1913, p. 6714. 



482 



COMPRESSED AIR PL.\XT 



may be a part of the air main, discomiected from the service 
hne so as to discharge into the atmosphere. On the receiver 
is an accurately cahbrated pressure gage. 

To begin a test, first open valves A and B \s-ide; then, with 
the compressor at normal speed, gradually close valve B until 
the tank pressure is constant. After a short run. to insure that 
pressure and temperature conditions have become constant, close 
valve A tight and allow the tank pressure to rise to lo or 15 lbs. 
above normal working pressure. Then close valve C, stop the 

70 



60 



50 



40 



530 



20 



10 



V 




\J 


k. 
































N 


X^^H^ 














N 

N 


'n^ 


"^ 


^ 


















K| 












> 


■\n 



50 



75 100 125 150 175 200 
Seconds 



Fig. 253. — Time-Pressure Cur\-e. 

compressor, open valve A quickly, and take exact simultaneous 
readings of time and pressure, as the tank pressure falls due to 
escape of air through valve B. An assistant holding a watch 
should call ''read" at uniform interA'als of lo or 15 seconds, con- 
tinuing until the pressure has dropped considerably below the 
working pressure. Finally, plot the readings on cross-section 
paper, to obtain a time-pressure curve (Fig. 253). 

From Chapter III, p. 50, if the weight of the volume of com- 

PV 



pressed air in the tank = IF. then F]' = ]VRT, whence Tr = 

W is taken as the entire volume of the tank and its connections 



Rr 



MEASUREMENT OF AIR CONSUMPTION 483 

between valves B and C; it is determined by weighing the volume 

of water required to fill the tank and connections, and dividing 

by 62.355 (weight of i cu. ft. of water at 62° F.). 

Differentiating the above equation with respect to time: 

dW V dP . ^. ^ dW . ^ , ^ , . , 

^— =--— X-r, m which -^- is the rate 01 decrease of weight 
dt Ri dt dt 

of air in the tank, or the rate of discharge in pounds per second, 

dP 
while — is the rate of decrease of pressure. 

0/1/ 

In Fig. 253, M is a point on the time-pressure curve corre- 
sponding to the normal working pressure. MN is carefully 
drawn tangent to the curve at M, and MK is parallel to the 

^, dP MK ^ , . . . , ,.,, 
pressure axis. Ihen —r=^=^r—. Substituting in the dmerential 

dt KN 

dW V MK 
equation: -— - = 7— X -7;^r7. Reading from the diagram the 

dt Kl KIS 

values of MK and KN, the rate of discharge, in pounds per 

second, is found directly. As the weight of air discharged 

equals the weight drawn into the compressor, the free air capac- 

WRTq 
ity is found trom Fo = — 5 — , where Vq is the volume, in cubic 

Pq 

feet of air per second, at the intake pressure Pq (pounds per 
square foot absolute) and at the absolute temperature Tq, and W 
is the weight of air discharged per second as already found. 

Another mode of using a single tank, especially adapted to 
testing rock-drills, is shown in Fig. 254. A vertical tank or 
receiver is connected to the main air line from the compressor 
by a I -in. pipe, another i-in. pipe leading to the rock-drill. On 
the receiver is a line of overlapping water-gages, or a single long 
gage glass, extending from bottom to top. Connected to the 
bottom of the receiver is a 3-in. pipe from a pump (or a water 
tank at an elevation sufficient to give a static pressure equal to 
the air pressure). 

To begin a test, run in enough water to show at the bottom of 
the gage glass, and make a chalk mark opposite this point. 
Next, turn on the compressed air until the receiver is filled at the 
desired pressure; then turn off the air and start the drill. During 



4^4 



COMPRESSED AIR PLANT 



the run, maintain a constant-pressure in the receiver, as indi 
cated on the pressure gage, by regulating the valve in the water 
pipe. When. the drill is stopped, mark instantly the height of 
water level on the gage glass by another chalk mark. Note 
the distance between the marks, and the elapsed drilling time. 
The cross-section of the receiver being known, compute the vol- 
ume of compressed air used, and the corresponding volume of 




Tee for 
Water Discharee 



Fig. 254. — Diagram of Tank for Measuring Air Consumption of Rock-Drills. 

free air. From the depth of hole drilled the air consumption per 
Hnear inch or foot of hole may also be found. 

If it be assumed that the receiver pressure is constant, the 
volume V of free air used is 

where R — rlse of water level, feet; A = cross-section of receiver, 
square feet; P = initial gage pressure, pounds; and -9" = atmos- 
pheric pressure. 



MEASUREMENT OF AIR CONSUMPTION 485 

In case the receiver pressure decreases from P to P' during 
the test, let F = initial volume of free air in the receiver at pres- 
sure P, and V = final volume of free air at pressure P' . Then 

F=^5(^)andF' = ^(£-ie)(^), 

where A, R and H are as above, and B = total height from initial 
water level to top of receiver, feet. Then the volume of free air 
used =F-F'. 



INDEX 



A 

Abrams, H. T., tests on air-lift pumps, 422 

Absolute pressure, temperature and zero, 47 

Adiabatic compression, 53, 56, 64, 138, 139, 140; equation of, 58, 65 

Adiabatic expansion, 241, 243 

Adjustable steam cutoff, valve 24 

Aff elder, W. L., on compressor explosion, 204 

Aftercoolers, 32, 96, 170, 173 

Ainsworth (B. C), hydraulic compressor, 213-215 

Air and steam cards combined, 23 

Air cards, 58, 97, 105, 146, 151-153, 189; elements of, 146, of stage compressor, 97 

Air-cataract valves, 121 

Air compression at altitudes above sea-level, 191-197; by direct action of falling 
water, 209-220 

Air compressors: belt-driven, 25, 26, 31, 33; builders, list of, 46; chain-driven, 20, 
37,40; classification, 8, 46; direct-driven, 38, 39, 41; dry, 76; for compressed- 
air haulage, 453-457; geared, 31, 38, 40; half-duplex, 20; horse-power of, 
138-143; hydraulic, 209-220; makers of, 46; performance of, 137 et seq.; 
tests, 149-168; water-driven, 26-33; wet, 72-75 

Air-consumption of machine drills, 295-301; testing for, 470, 471, 478-485 

Air cylinder governors and unloaders, 182-190 (see Air-pressure regulators) 

Air cylinders, proportions of, 25, 88-92, 95 

Air, discharge of, through circular orifices, 469 

"Air-electric" drill, 5, 290-292; track channeler, 389, 392 

Air engines, 235-252 

Air-feed hammer drills, 319, 331, ^Z2>^ 339-346 

Air, flow of, through orifices or short tubes, 467-469; through pipes, 221; under 
small pressure, 468 

Air, free, 47; specific heats at constant pressure and volume, 53; weight of, 48, 50, 
469 

Air governors, 175-190 

Air hammer drills, 311 e/ seq. 

Air inlet, area of, 100, 123; conduit for, 117; valves, 99-118 

Air-lift pumps, 415 et seq.; air pressure for, 418, 419, 420, 421, 424, 426; efficiency, 
419,420,421,426,431; examples of, 424-431; foot-pieces, 422-425; ratio of 
submergence to lift, 421; tests, 421-426; volume of air for, 419; working 
under heavy heads, 431 

Air, measurement of flow of, 467-485 

Air pressure for machine drills, 295-297 

487 



488 INDEX 

Air-pressure regulators, 175-190; Allis-Chalmers, 180, 185; American, 177; 
Clayton, 176; Franklin, 180; IngersolL-Rand, 177, 180-185, IQOJ Laidlaw- 
Dunn-Gordon, 180, 181; Nordberg, 180, 185-189; Norwalk, 177; SuUivan, 
178, 179, 180; unloaders, 182-190. 

Air pulsator for "electric-air" drills, 290 

Air receiver, arranged for measuring compressed air, 481 

Air receivers, 96, 169 et seq.; baffle plates for, 173; functions, 169-171; sizes, 169, 
174; underground receivers, 171, 172, 256, 406; volumetric capacity, 174 

Air- thrown valves for rock drills, 274-288 

"Air-tube Rotator" drill (Sullivan Machinery Co.), 328, 329 

Alaska-Treadwell Gold jMining Co., compressor plant, 26 

Alley & MacClellan compressors, 8, 38 

Allis-Chalmers compressors, 10, 19, 46, 78, 99; mechanically-controlled valve- 
motions, 129; plate valves, no; pressure regulators and governors, 180, 185 

Altitudes above sea-level, air compression at, 191 et seq.- barometric pressures 
corresponding to, 194; mechanically-controlled inlet valves for, 195; stage 
compression at, 196 

American Air Compressor Works, 46; air governor, 177 

American Inst, of Mining Engineers, Trans, of, 143, 205, 225, 249, 251, 427, 429, 463 

American Locomotive Co., compressed-air locomotives, 436 

American Machinist, 8, 192, 246, 267 

American Society of Civil Engineers, Proceedings of, 415 

Anaconda copper mine, compressed-air haulage at, 461, 462; compressed-air 
hoists, 249-251 

Angelo and Cason Mills, So. Africa, tests on air-lift pumps, 424-426 

Angle-compound compressor (Sullivan Machinery Co.), 37, 117 

Ante-coolers for compressors, 96 

Aragon iron mine, Mich., compressed-air haulage at, 462 

Area of compressed air inlet, 100, loi; of discharge, 122, 123 

Area of discharge valves, 122 

Association of Engineering Societies, Transactions of, 406 

Auger coal drills: Fairmont Mining Machinery Co., 377; Ho wells, 377; Ingersoll- 
Rand, 376; Jeffrey, 377 

"Auger Rotator" drill, Sullivan Machinery Co., 329, 331 

Au.xiliary reservoirs for compressed-air locomoti\es, 445, 446, 447, 459, 463 

B 

Baffle plates for air receivers, 173 

Bailey & Co., Manchester, compressor piston valve, 136 

Bailey Meter Co., 476 

Balanced piston valves, 24 

Baldwin Locomotive Works, compressed-air locomotives, 435-437, 444 

Ball-and-socket joints for compressed-air locomotive charging station, 449, 450, 

451 

Beckett, P. G., on air-lift pump, 427 

Behr, H. C, air-lift pump experiments, 418; on "two-pipe" system of com- 
pressed air transmission, 247 

Bell, J. E., experiments on rock drills, 299 



INDEX 489 

Bellis & Morcom compressors, 8 

Belt-driven compressors, 20, 25, 7^^, 35-39 

Bendigo district, Victoria, Lansell's air-lift pump, 427, 428 

Bends in air pipe, 233, 234 

Bernstein, P., on Clausthal hydraulic compressor, 219 

Berwind-White Coal and Coke Co., compressor tests, 151, 152 

"Best receiver pressure" for compressors, 60, 62, 67 ' 

Bit-shanks for drills, 315, 318, 335, 337, 344 

Bits, hollow, 313, 315, 320, 322, 337, 347; solid, 327, 330, 344, 346, 347 

Blakney, hydraulic compressor, 219 

Bleeder valve for compressed-air locomoti\'e charging-station, 450 

Bowden, J. H., test on compressed-air haulage plant, 463 

Boyle's law, 48, 49, 50 

Brandt rotary drill, 269 

Breakage of drill parts, 305 et scq. 

British- Westinghouse Rateau mixed-pressure turbine and air compressor, 42, 43, 45 

Brotherhood & Co., Peter, compressors, 8 

Bryant channeler, 384 

Buck Mountain Colliery, compressed-air haulage at, 458 

"Bullmoose" hammer drill, 325, 327 

Burbanks Main-lode mine, W. Aust, two-stage air-lift, 430, 431 

Burleigh compressor at Hoosac Tunnel, 2 

Burning-points of cylinder oils (see Ignition points) 

Burra-Burra mine, compressor at, 21, 149 

Bury Compressor Co., 46 

Butte, Montana, compressor explosion, 204 

"Butterfly-valve" drills, Ingersoll Rand Co.: hammer, 313-316, 323-327, 343- 

346; reciprocating, 283-285 
By-pass for air cylinder, 82 

C 

Cable-reel electric locomotive, 132 

Calumet and Hecla Copper Mine, introduction of machine drills, 2 
Canadian Electrical Neivs, 213 
Canadian Engineer, The, 210 
Canadian Mining Institute, Transactions of, 256 
Capacity of air for moisture, 84, 253 
Carnahan, C. T., Manufacturing Co., hammer drill, 349 
Carriage mounting for drills (in tunneling), 273 

Cason and Angelo mills. South Africa, tests on air-lift pumps, 424, 426 
Cataract valves for compressors, 121 
Cave rock drill, i 

Chain coal-cutters: Jeffrey, 350, 352-357; Sullivan, 350, 351, 357-361 
Chain-driven compressors, 20, 37, 40 
Chamber of Mines, W. Aust., Jour, of, 431 
Champion Iron mine, Mich., drill tests, 300 

Channeling machines, 380 et seq.; depth of cut and speed of work, 388; shape of 
bits, 380, 385, 389 



490 INDEX 

C banning, J. Parke, air-compressor test, 21, 149 

Charcas, S. L. P., Mex., air-lift pumps at, 431 

Charging compressors for compressed-air locomotives, 453-457 

Charging stations for compressed-air locomotives, 449-451 

Charles's law, 49 

Chattering of inlet valves, loi^ 103 

Chersen drill, 287 

"Chicago Catling" drill, 286 

"Chicago Hummer" drills, 286, 331, 332, 334 

Chicago Pneumatic Tool Co., 287, 331, 335, 346; compressor, 27, no; hammer 
drills, 331, 334, 335, 346; reciprocating drills, 286, 287; underground com- 
pressed-air hoist, 249 

Chodzko, A. E., on "two-pipe" S3'stem of air transmission, 247 

Choking of air pipes by ice, 172 

"Choking" controller for air cylinders, 185 

Christensen compressor, 40 

"Cincinnati" air-valve gear, 129 

Clack \aives for air inlet, 99 

Clark. D. K., volumes and weights of air at different temperatures, 48 

Clark's Rules and Tables, 468 

Classifications: compressors, 8, 46; hammer drills, 311; reciprocating drills, 273 

Clausthal silver mines, hydraulic compressor at, 219 

Clayton compressor, 2, 46; governor and pressure regulator, 176 

Clearance: in air engine, 243 ; in compressor cylinder, 63-71, 79-82; proportionate 
and disproportionate, theory of, 67-70 

Clearance controller, for compressors, 185 

Clemens, G., data on compressed-air haulage, 459 

Cleveland hammer drill, 349 

Clifton Colliery, England, compressor explosion, 201 

Climax "Imperial" drills, R. Stcj^hens & Son: hammer t>'pe, 318, 319; recipro- 
cating type, 280-2S2 

Coal-cutting machines, 350-379; classification, 350; comparison of, 377; disk or 
circular saw, 362-365; depth and width of cut, 356, 361, 366, 371; endless 
chain, 350-359; pick or reciprocating, 365-374; rotary-bar, 359-362 (see 
Coal punchers) 

Coal picks (see Coal punchers) 

Coal punchers: Hardy, 371; Harrison, 366, 367; IngersoU-Rand, 369, 370; Pneum- 
electric, 290, 372-374; "Radialaxe," 371, 372; Sullivan, 365, 366, 368, 372 

Cobalt, Ont., hydraulic compressor at, 217 

Cochise rock-drills: hammer, 346, 349; reciprocating, 287 

Colladon, early compressors for rock-drills, i, 74 

Colliery Guardian, 256 

Colorado Fuel Co., use of Stanley Header by, 376 

Column mounting for drills, 272-273 

Combined air and steam indicator cards, 23 

Compaiiia de Penoles, electric-driven compressor, 38 

Comparison of types of compressors, 20-22 

Complete expansion of air, working with, 240, 241 



INDEX 491 

Compound air-driven pumps, 405-408 

Compound compressed-air locomotives, 433 et seq.] interheater for, 435, 436 

Compound steam end for compressors, 8, 13 et seq. 

Compressed-air drills, 269-349 

Compressed-air engines, 235-252; actual work done, 242; clearance, 243; con- 
sumption of air, 245; cutoff, 243, 244; cylinder volume for given horse-power, 
244; reheating for, 245, 248, 251; working at full pressure, 239; working 
with complete expansion, 240, 241; working with partial expansion, 
240, 248 

Compressed-air haulage, 432-466; charging stations, 449; locomotives, 433-447; 
pipe-lines, 447-449; plants, examples of, 458-466 

Compressed-air hoists, 247-252; air storage for, 250, 251; at Anaconda mine, 249- 
.251; at Copper Queen mine, 252; at Franklin Junior mine, 252; at Miami 
mine, 248, 252; at Simmer Deep mine, Witwatersrand, 247; reheating for, 
248, 251; volume of air for, 248, 250, 251, 252 

Compressed-air hoists (small) for underground use, 248-249; Chicago Pneumatic 
Tool Co.'s hoist, 249; Hohman hoist, 249; Ingersoll-Rand "Little Tugger," 
249; Leadville "column" hoist, 249 

Compressed Air Machinery Co., compressors, 31, 46 

Compressed Air Magazine, 116, 174, 218, 224, 244, 246, 258, 398, 399, 411, 426, 

459, 481 
Compressed air, measurement of flow of, 467-485 
Compressed-air pumps, 394-431; adjustment of air pressure, 397; efficiencies of, 

462; interheaters for, 406, 407; preheating for, 405; prevention of freezing 

in, 405, 407 
Compressed air, reheating of, 258-268; power cost of, 150 
Compressed-air transmission, 221-234; loss of power, 221; loss of pressure, 222; 

piping, 223, 225, 230, 231, 233; resistance due to bends in piping, 233, 234 
Compressed air versus electric transmission, 5 ; verstts steam transmission, 3 
Compressed air versus steam for direct-acting pumps, 397, 398 
Compression curve, construction of, 144 
Compression of air, 47 et seq.; adiabatic, 53, 56, 64, 138, 139, 141; at altitudes 

above sea-level, 191-197; by direct action of falling water, 209-220; dry 

versus wet, 82; heat of, 53, 54, 59, 76; isothermal, 53, 55, 63, 138, 139, 141; 

laws, 48, 49; rate of increase of temperature, 50, 51 ; stage compression, 60-64, 

66-71; temperatures of (see Temperatures); with clearance, 63-71; without 

clearance, 54-63 
Compressor, air valves, 99-136; explosions, 198-208; half-duplex, 20; indicator 

cards, 97, 105, 146, 151-153; piston speed, 32, 73, 77, 86, 95, loi, 123; tests, 

21, 45, 149-168 
Compressor builders, list of, 46 
Compressors: chain-driven, 20, 37, 40; classifications, 8, 46; comparison of types, 

20-22; Corliss steam end, 14, 17, 21, 23; direct-driven and belt-driven, 25, 26, 

31, 33 et seq.; dry, 76, 82; duplex, S, 15 et seq.; electric-driven, 20, 38, 40, 41; 

half-duplex, 20; high-speed, 41; horse-power, of 138-143; for locomotive 

haulage, 453-457; performance of, 137-168; proportions of cylinders, 25, 90, 

92, 95; rating of, 137, 138; stage, 8, 10, 13-19, 86; straight-line, 8, 9 et seq.; 

turbo-compressors, 42-45; vertical, 8; water-driven, 26-33; wet, 72, 82; 



492 INDEX 

with compound air end, 8, lo, 13-19, 86; with compound steam end, 8, 13 
et seq. 

tompressors, makers and names of: Alley & MacClellan, 8, 38; Allis-Chalmers, 
10, 19, 46, 78, 79; Bellis & Morcom, 8; Burleigh, 2; Bury, 46; Chicago 
Pneumatic Tool Co., ay-, 46; Christensen, 40; Clayton, 2, 46; Compressed 
Air Machinery Co., 31, 46; Dubois-Franfois, 2, 72, 99; Frankhn, 180; Hanarte, 
73; Humboldt, 72, 73; Ingersoll-Rand, 10, 11, 15, 16, 17, 24^ 26, 28, 39, 40, 
42, 45, 46, 80, 81, 99, loi, 152, 453, 455-457; King-Riedler, 8; Laidlaw-Dunn- 
Gordon, 9, 10, 17, 18, 36, 46, 79, 81, 99, loi, iiOj 202; Leyner, no, in; 
Nordberg, 21, 30, 46, 76, 77, 99^ 149; Notwalk, 2, 10, 13, 45, 46, 89, 99, io6j 
125, 452; Rand, 2; Rateau (turbo), 42; Riedler, 8, 20, 132; Rix, Com- 
pressor and Drill Co., 46; Robey & Co., 8, 20, 41; Schneider & Co., Creusot, 
France, 99; Sullivan Machinery Co., 10, 12, 14, 34, 35, 37, 46, 99; Vulcan 
Iron Works Co., 66; Walker Bros., 20; Worthington Pump and Machine 
Co., 46 ■ 

Conduit for inlet air, 11 7-1 18 

Consumption of air: by air engines, 244, 245; by direct-acting pumps, 399-402; 
by machine drills, 285, 286, 287, 296, 301, 347; by pneumatic displacement 
pumps, 411, 415 

Controller, clearance, for compressors, 190 

Conveyance of compressed air in pipes, 221-234 

Coolgardie, Western Aust., air-lift at, 430, 431 

Cooling: aftercoolers, 32, 96, 170, 173; ante-coolers, 96; modes of, 59, 60, 64, 
72-75, 76-79, 82-85, 87; intercoolers, 32, 60, 62, 67, 87, 90, 93-96; in receivers, 
60, 95, 97, 174; water-jackets, 59, 76-79 

Copper Queen mine, compressed air-hoist, 252 

Corbett, R. H., on compressed-air hoist, 252 

Corliss air valves, 10, 14, 17, 19, 21, 32, 99, 107, 121, 124, 129, 130, 131 

Corliss steam-valve motion for compressors, 14, 17, 21, 23, 99 

Costs of air compression, 155, 157, 158, 159, 161, 162, 164, 165, 166, 217, 218, 219 

Couch, J. J., rock-drill, i 

Cox, Wm., compressed-air calculations and tables, 399-401 

Crispell, C. W., nomograms for compressed-air calculations, 142, 143, 225 

Cummings, Chas., "two-pipe" system of compressed-air transmission, 245, 247, 

408 e«j 

Cushioning in machine drills, 276, 293, 303, 304, 316 

Cutoff in air engines, 243, 244, 248; nominal and actual, 242, 243 

Cylinder dimensions of simple pumps, 398 

Cylinder oils, flash and ignition points, 200, 202, 204 

Cylinder proportions for compressors, 25, 90-92, 95 

Cylinder volumes: in stage compression, 90-92; of air engin2, 244, 245 



"Dancing" or "chattering" of inlet valves, loi, 103 

D'Arcy's formula for loss of pressure in pipes, 223, 230; graphic solution of formula 

225, 228, 229 
De Kalb, 111., tests on air-lift pump, 418 



INDEX 49^ 

Delivery valves, 119-123, 124-136; cataract-controlled poppets, 121; effect' of 
leakage, 119, 201; mechanical control for, 125; poppet type, 119, 127, 129 

Denver Rock Drill and Machinery Co., hammer drills, 335-337, 340-342 

Deposition of moisture from compressed air, 170, 1 71-173 

Depth of hole: hammer drills, 347, 348; reciprocating drills, 309 

Dickson Manufacturing Co., compressed-air locomotives, 461 

Direct-acting pumps ,operation by compressed air, 394-408 

Direct compression by falling water, 209-220 

Direct-connected compressors, 38, 41 

Disk or circular saw coal cutters: Jeffrey, 362, 364; Mavor & Goulson (electric 
only), 7, 365 

Discharge valves (see Delivery valves) 

Discharge-valve area, 122 

Displacement pumps, pneumatic, 409-415; Halsey, 412; Latta-Martin, 411; 
Merrill, 410 

Double-ported inlet valves, 129 

Drill repairs, 304; costs of, 305-307 

Drilling records, 309, 348 

Drills, rock: air pressure for, 295 et seq.; hammer drills, 311 e/ seq.; mountings, 
270-273; reciprocating drills, 273 et seq.; repairs, 304-308; records of work, 
300, 301, 309, 348; "stoper" drills, 339; tables of drilling speeds, 309, 348; 
valveless drills, 293, 320, 327, 328 

Drummond Colliery, compressed-air pumps at, 256 

Dry compressors, 76-85 - 

*'Dry" reheaters, 268 

Dry versus wet compression, 82 

Dubois-Frangois wet compressor, 2, 72, 99 

Duplex compressors, 8, 15-20, 21, 29, 30, 36-39, 41 

Dust, cause of Silicosis, 316; effects on spool-valve drills, 303 

Dust allayers for machine drills, 282, 319, 346 

Duty of compressors (see Output of compressors) • 

Duty of machine drills, 300, 301, 309, 348 

East Rand Proprietary Mines, Ltd., tests on air-lift pumps, 424-426 
Ebervale, Luzerne Co., Pa., tunnel at, 231 

Efficiencies of air-lift pumps, 419, 420, 421, 426, 431; of direct-acting compressed- 
air pumps, 402, 403, 404, 408 
Efficiency of compressors, 137-168 (see also. Output of compressors) 
Efficiency of reheating, 259, 260, 261, 262, 267, 268 
Efficiency of rock drills, 301, 304 
Electric-driven compressors, 20, 38, 40, 41, 250 
Electric-driven rock drills, not satisfactory, 6 
"Electric-air" drill (see "Air-electric" drill) 
"Electric-air" track channeler, 389, 392 
Electric versus compressed-air haulage for mines, 432 
Empire mine, Grass Valley, Cal., compressed-air haulage at, 458, 459 



494 INDEX 

Endless-chain coal cutters, 350-359 

Engineer, The (London), 424 

Engineering and Mining Journal, 210, 213, 216, 218, 219, 249, 252, 273, 295, 472, 

473. 475 
Engineering News, 81, 200, 418, 419 
"Engineering Thermodynamics," Lucke, 60, 67 
Expansion curves, air and steam, 236, 237 
Explosions in air compressors and receivers, 198, 208; examples, 201-205; P^^' 

cautions for preventing, 206 
Externally heated or "dry" reheaters, 261-264 



Fairmont Mining Machinery Co., auger drills, 377 

"Feather" valve, Laidlaw-Dunn-Gordon Co., 9, 18, 21, 36, no, 113 

Federated Instn. of Min. Engs., 199, 201 

Fergie, Chas., mode of preventing freezing of moisture, 256 

Fessenden, E. E., 481 

Field of work, hammer drills, 348, 349 

Final temperature of air compression (see Temperatures of compression) 

"Fitchering" of drill holes, 269, 299, 304 

Flash and ignition points of cylinder oils, 200, 202, 204 

Fleigner's formula for flow of air, 473 

Flottmann hammer drill, 349 

Flow of air through orifices or short tubes, 467-469 • 

Flyball governors, 176, 177, 179, 181 

Foot-pieces for air-lifts, 422-425 

Four-stage compressor, 453, 456, 457 

Fowle, rock drill, i 

FrankHn compressor, 180; pressure regulator and unloader, 180 

Franklin Junior copper mine, compressed-air hoist, 252 

Fraser & Chalmers, Riedler compressor, 20, 132 

"Free" air, 47 

Freezing of moisture in compressed air, 84, 172, 253-257, 404, 407; deposition of 
moisture by reducing pressure, 256; influence of high pressures in transmis- 
sion, 255; protection of surface piping, 256 

Frick, H. C , Coal and Coke Co., compressor explosion, 202; use of Stanley header, 
376 

Frictional losses in compressors, 21, 85, 137, 140, 150, 152, 156, 165 

Frictional resistance in air pipes, 222 et seq.; due to bends, 234 

Friedrich, G. C. H., tests on air-Hft pumps, 421 

Frizell, J. P., on hydraulic air compression, 209 

Fuel cost of reheating, 258, 259, 260, 261, 262, 272 

Full pressure in air engines, working with, 239 

Functions of air receiver, 169-171 

G 

Gasolene-driven compressors, 25, 26, 27 
"Gathering" locomotive, electric, 432 



INDEX 495 

\ 

"Catling" drill, Chic. Pneum. Tool Co., 286 

Gay-Lussac's law, 49 

Geared compressors, 31, 38, 40 

Gibson-TngersoU "Electric-air" channeler, 389, 392 

Gilman, G. H., apparatus for measuring compressed air, 481 

Glen Lyon, Pa., Colliery, compressed-air haulage at, 462, 463 

Goleta Mining Co., water-driven compressor, 29 

Goodman Mfg. Co., electric coal-cutters, 350 

Governors, air: American, 177; Clayton, 176; Franklin, 180; Inger soli-Rand, 

177,180-185,190; Laidlaw-Dunn-Gordon, 180,181; Nordberg, 180,185-189; 

Norwalk, 177; SuUlivan, 178, 179, 180 
GOw, A. M., flashing and ignition points of lubricating oils, 200 
Grades of mine haulage tracks, 452, 453, 458, 459, 461, 462, 463 
Graphic determination of horse-power of compressors, 142 
Guttermuth, experiments on reheating, 262; plate valve, no 
Gwin mine, Cal., reheater for pump, 407 

H 

Half-duplex compressor, 20 

Halsey, F. A., compressed-air data, 97, 192, 193, 230; pneumatic displacement 

pump, 412 
Hammer drills, 311-349; classification, 311 ; depth of hole and speed of work, 34S: 

makers of, 349 
Hanarte wet compressor, 73 / 

Hardsocg Wonder Drill Co., 349; hammer drills, 312, 320, 321 
Hardy coal-pick, 371 

Harris, Elmo G., on pneumatic displacement pumps, 415 
Harrison pick machine, 366, 367 
Haulage by compressed-air locomotives, 432 et seq. 
Heat curves, 50-53 
Heat, losses in compressors, 53, 59, 64, 83, 137; transference (abstraction) of, 53, 

59 
Heat of compression, 50-53, 76; absorption of, 59 
Heating of air-cylinder walls, 73, 79 
Henderson and Wilson, tests on air-lift pumps, 424-426 
High altitudes: mechanically-controlled inlet valves for, 195; stage compression at, 

196; work of compressors at, 191-195 
High-pressure air, measurement of, 471 

High-pressure transmission of air, as influencing freezing, 246 
"High-range" compressed-air transmission (see "Two-pipe" system) 
High-speed compressors, 41 
Hirschberg, C. A., compressor data, 95 
Hiscox, G. D., tables of compressed-air engine data, 242, 243 
Hitchcock, C. K., Jr., on rock-drill repairs, 305-307 
Hoisting engines, compressed-air, 247-252 
Holdsworth, F. D., device for measuring compressed air, 472 
Holman drills: spool- valve, 282, 283; tappet, 290 
Holman underground compressed-air hoist, 249 



496 INDEX 

Homestake gold mine, compressed-air locomotives at, 433, 462 

Hood, B. B., 475 

Hoosac tunnel, Burleigh compressors at, 2 

Horse-power: of air engines, 242-244; of air-lift pump, 420, 421, 422; of com- 
pressors, 95, 138-143, 150, 151, 155-168; graphic determination of, 143 
^Horse-power per cubic foot of free air, 139-141 ; allowance for losses in compression, 
139, 140, 141 ; at high altitudes, 194; for any number of drills at any altitude, 
298; modifying factors, 298, 299; theoretical, 138, 141, 147-148 

Howells Mining Drill Co., auger drill, 377 

Hiimboldt wet compressor, 72, 73; rubber-ring valve, 72 

Humidity of atmospheric air, 84 

"Hummer" drills, Chicago Pneumatic Tool Co., 286, 331, ^3^, 334 

"Hurricane-inlet" valve, 107-109, 118 

Hydraulic air compressors, 209-220; Blakney, 219; at Clausthal, 219; at Cobalt, 
Ont., 217; at Kootenay, B. C, 213-215; ^SlacFarlane, 219; Magog plant, 
209-213; at Norwich, Conn., 218; at Petersborough, Ont., 218; at Victoria 
mine, 215; Washington, 218 

"Hy-speed" drill, Sullivan Machinery Co., 280 



Ignition points of lubricating oils, 200, 202, 204 

Her Rock Drill Manufacturing Co., hammer drill, 349 

"Imperial" compressors, Ingersoll-Rand Co., 15, 16, 17, 40 

Indicator cards, air, 58, 97, 105, 146, 151, 189; air and steam cards compared, 2^ 

Ingersoll-Rand Co.: after-cooler, 170, 173; air controller, 190; "air-electric" 
drill, 290-292; air-lift pump, 422, 424; auger drill, 367; "Butterfly" valve 
drill, 283-285, 313-316, 322, 325, 343, 344; channelers, 382, 384, 386, 387, 389, 
390, 392; coal pick, 369; compressors, 10, 11, 15, 16, 17, 24, 26, 28, 39, 40, 42, 
45, 46, 80, 99, loi, 152, 453, 455-457; foot pieces for air-lift pumps, 422-425; 
hammer drills, 313-317, 322-327, 343-345; "Hurricane-inlet" valve, 107- 109, 
118; intercoolers, 94, 96; pick machine, 369; piston-inlet valve, 107; plate 
valves, 10, 114; pressure regulators, 180, 182-185; "Radialaxe" coal cutter, 
371; ram track channeler, 382, 384; receivers, 96, 190; receiver aftercooler, 
96, 173; reciprocating drills, 274-277, 283-286, 301; reheater, 263; "return- 
air" displacement pump, 413-415; Sergeant drill, 274-277, 292; speed and 
pressure regulators, 177, 180, 190; stoper (air-feed) drills, 343 -346; Temple- 
Ingersoll "electric-air" drill, 290-292; underground compressed-air hoist, 
249; air unloaders, 182-185, 190 

Injection water for compressors: effects of, on air cylinder, 83, 85; quantity of, 
42, 75; temperature of, 33,, 74, 75, 97 

Inlet air, arrangement for admitting, 116-118; temperature of, 117 

Inlet area of air cylinders, 100 

Inlet conduit for air cylinders, 117 

Inlet \alves, 99-118; area of, 100; "chattering" of, 103; requisites of, 99 

Inspiration Consol. Copper Co., Ariz., compressed-air haulage, 465-466 

Institution of Mining Engineers (England), Trans, of, 199, 201 

Institution of Civil Engineers (London), Proceedings of, 209 



INDEX 497 

Intercoolers, 32, 60, 62, 67, 87, 90, 93-96, 97; IngersoU-Rand, 94, 96; NorwalL, 

89; Sullivan, 96 
Interheater: for compound compressed-air locomotives, 444, 447; for compound 

pumps, 405-407 
Internally fired reheaters, 264 

Isothermal compression, 53, 54, 63; equation of, 53, 55, 64 
Ivens, E. M., book on "Pumping by Compressed Air," 409 
Izod, E. G., drill tests, 300 

J 

"Jackhamer" drills, IngersoU-Rand Co.: "BuUmoose," 325; hand drills, dry and 

wet, 322-324, 325, 326; mountings for, 324;: "Sinker," 326 
Jeddo (Pa.) mining tunnel, compressed-air transmission in, 231 
Jeffrey Mfg. Co., coal cutters, 350, 351, 352-357 

Jeffrey coal cutters: "longwall" machine, 351-355; "shortwall," 356-357 
Jeffrey loading machine, 378, 379 

Johnson, E. E., on performance of air-lift pump, 419, 420 
Joule's heat unit, 51, 239 

K 

"K," values of, 54 ' 

Kennedy, Alex. B. W., reheating tests by, 261 

Kent, Mech. Eng. Pocketbook, 145 

Kerosene, use of, in compressors, 204, 207 

King-Riedler compressor, 8 

Knight water-wheel, 26 

Kootenay (B. C), hydraulic air compressor at, 213-215 ; 

Koster piston air valve, 136 , 



Laidlaw-Dunn-Gordon Co., 46; air governor, 180, 181; compressors, 9, 10, 17, 18, 
36, 46, 79, 81, 99, loi, no, 202; "Cincinnati" valve gear, 129; "feather" 
valve, no, 113; mechanically-controlled valve motions, 128, 129; poppet 
valves, 102, 120; pressure regulator, 180, 181 

Lansell's air-lift pump for shafts, 427, 428 

Laschinger, E. J., drill tests, 300 

Latta-Martin pneumatic displacement pump, 411, 412 

Laws governing compression of air, 47 et seq. 

Leadville, "column" hoist, 249 

Leakage of air valves, 105, 119 

Leakage of compressed-air pipe lines, 231, 233, 449, 463 

Leaky air piston, effect of, 98 

Leaky discharge valves, 119 

Lees, T. G., on compressor explosions, etc., 78, 199, 201 

Leyner compressor, no; plate-valve, 110-112, 115 

Leyner-Ingersoll water drills, 286, 313-316 

"Liteweight" drill, Sullivan Machinery Co., 277, 278 

"Little Champion" drill, 289 



498 INDEX 

"Little Tugger" compressed-air hoist, 249 

Loading machines for working places underground, 379 

Locomotives, compressed air, 432-466; construction and operation, 433 et seq.; 

cylinder pressures, 434, 435, 436, 440, 441, 442, 445, 448, 453, 458 et seq. 
Locomotive charging compressors, 453-457 
Longwell coal cutters, 351, 352, 357, 359, 363, 364 
Loss of head in pipe transmission, 222, 226, 227, 230; of power, 221 
Loss of volumetric capacity due to piston clearance, 79-82 
Losses in air compression, 137-140, 194 

Losses in transmission of compressed air, 221-223, 226, 227, 230-233 
Low-pressure air, measurement of, 469 
Lubricating oils, flash and ignition points, 200, 202, 204 
Lubrication of air cylinders, 198, 200, 202, 203, 204, 205, 206; of rock drills, 316, 

327, 328, 344 
Lubrication, "flood and splash" system, 40; quantity of, 206; use of soap and 

water, 206 
Lubricators, sight-feed, for compressors, 206 
Lucke, C. E., Engineering Thermodynamics, 60, 67 



M 

Machine drills, 269 et seq.; air pressures for, 295; classifications, 273; column and 
bar mountings, 272, 273; consumption of air by, 295-301; cushioning of 
stroke, 288, 303, 304; dust allayer for, 281, 282, 319; efficiency, 301-304; 
"electric-air" drill, 290-292; foul air from, 205; general description, 269; 
hammer drills, 311; length of stroke, 269, 303; modes of mounting, 270-273, 
324, 332, T,7,y, records of work, 287, 296, 300, 301, 309,.3io, 348; repair costs, 
305-308; repairs of, 304-308; sizes (see different makes of drills); speeds of 
drilUng, 296, 300, 309, 348; speed of stroke, 269, 285, 292, 301, 304; spool 
and other air-thrown valve drills, 274-288; tappet- valve drills, 288-290; 
testing for air consumption, 470, 471, 478-485 

Machine drills (hammer type), classification, 311; makes: Cleveland, 349; "Climax 
Imperial" (Stephens'), 318; Cochise, 346; Flottmann, 349; "Chicago Hum- 
mer," 331, 335; Hardsocg, 320; Her, 349; IngersoU-Rand, "Imperial," 
327; "Jackhamer," 322-327; Leyner-IngersoU, 212-316; Murphy, 322; 
Shaw, 349; Sullivan, 316-318, 328; Waugh, 335, 337, 340-342; Whitcomb, 
349; "Wonder" (Hardsocg), 320, 321; Wood, 339 

Machine drills (reciprocating), classifications, 273; makes: Chersen, 287; "Chi- 
cago Catling," 286; "Chicago Hummer," 286, 331, 334; "Chicago Slugger," 
287; "Climax Imperial" (Stephens'), 280; Cochise, 287; Holman, 282; 
IngersoU-Rand "Butterfly," 283; McKiernan-Terry, 287; Murphy, 289; 
"Sergeant," 274; Siskol, 287; Sullivan, 278; Temple-IngersolU 290; Tri- 
umph, 293; Wood, 279 

Magog (Prov. Quebec), hydraulic air compressor, 209; reheating at, 267 

Mallard, M., working of compressed-air engines, 238 

Mavor & Coulson coal-cutters, 361-363 

McFarlane hydrauUc air compressor, 219 



INDEX 499 

McKiernan-Terry Drill Co., 287; hammer drills, S37} 33^^ 3395 reciprocating 
drill, 287 

McLeod, C. H., tests on Magog hydraulic compressor, 213 

Mean pressures in air compression, 144, 145, 194 

Measurement of compressed air: by meters, 476-477; through nozzles, 473; 
through orifices, 472; through short tubes, 467; by pitot tubes, 475; by 
tanks, 470, 471, 478-485 

Measuring tanks for compressed air: single tank, 481-485; with pair of tanks, 
470, 471, 478 

Mechanical Engineer's Assoc, of the Witwatersrand, Trans, of, 247, 295 

Mechanical Engineer's Pocketbook, Kent, 145 

Mechanically controlled air valves, 99, 107, 124-136; Allis-Chalmers, 129; Amer- 
ican, 132; Clayton, 132; disadvantages for delivery valves, 125; for high 
altitudes, 195; Franklin, 132; Koster, 136; Laidlaw-Dunn-Gordon, 128; 
Nordberg, 128; Norwalk, 125; Riedler, 132-136; Rix, 132; Sullivan, 131 

Merrill pneumatic displacement pump, 410-41 1 

Meters for measuring air, 476-477 

Meyer steam-valve gear, 10, 23, 156 

Miami copper mine, compressed air hoist, 252 

Midland Inst, of Min., Civ. and Mech. Engs., Trans., 38 

Miner's consumption (Sihcosis), 316 

Mines and Minerals, 204, 216, 218, 295 

Mining and Scientific Press, 249, 305, 478, 479 

Modern Machinery, 247 

Moist air, effect of, in compression, 83, 85 

Moisture in air, 83, 84, 172, 253; freezing of, 84, 172, 253, 257 

Mont Cenis tunnel, 1,2; speed of advance in, 2 

Morgan-Gardner Electric Co., coal-cutters, 350 

Morison, J., flashing-points of lubricating oils, 200 

Morning mine, Mullan, Idaho, compressor plant, 32-33 

Mule haulage, cost of, 458, 461, 464 

Murphy hammer drill (C. T. Carnahan Co.), 322 

Murphy "Little Champion" reciprocating drill (Wickes Bros.), 289-290 

Mushroom inlet valves, loi, 107 

Mylan, W. F., compressor tests, 45 

N 

"m," values of, 54, 60, 77, 138 

Napierian logarithm, value of, 55, 138 

New Hucknall colliery, turbo-compressor, 44, 45 

New Jersey Meter Co., 476 

New York Air Compressor Co., 46 

New York Aqueduct, explosion in compressor, 204 

Nicholson, J. T., reheating experiments by, 267 

Nominal and actual cutoffs in air engines, 242-244 

Non-conducting covering for air pipe, 4 

Nordberg, B. V., Anaconda compressed-air hoists, 249 

Nordberg compressed-air hoist, Miami mine, 252 



500 INDEX 

Nordberg^ compressor, 21, 30, 46, 76, 77, 99, 185; air-pressure regulator, 180, 
185-189; constant speed, variable delivery compressor, 188, 189 

Norris, R. V., tests on compressed-air haulage plant, 463 

North Star mine, reheating at, 267 

Norwalk compressor, 2, 10, 13, 45, 46, 89, 99, 125, 453, 454, 458, 463; intercooler, 
89; poppet inlet valve, 102, 106; "skip-valve," 106; pressure regulator, 177; 
receiver, 169 

Norwalk Iron Works Co., 10, 232, 234 

Norwich (Conn.), hydraulic air compressor, 218 

Nozzles for measuring flow of air, 473-475 

O 

Ohio Society of Mech., Elec., and Steam Engineers, Trans, of, 421 
Oil-cataract delivery \alves, 121 
Oiling devices for machine drills, 316, 327, 344 

Oils, lubricating, 200, 206, 207; flash and ignition points, 200, 202, 204 
Old Dominion copper mine, air lifts at, 427, 429 
Orifices, measurement of air through, 467-469, 472-475 

Output of compressors, 21, 22, 26, 43, 44, 45, 83, 95, 97, 105, 138-143, 154, 156- 
160, 162-165, 168, 195 



Partial or incomplete expansion in air engines, 240 

Pauly, K. A., 249 

Peele's Mining Engineer's Handbook, 215, 316, 421 

Peerless Colliery (W. Va.), compressed-air haulage at, 459 

Pelton water-wheel, 26, 30, 31, ;^2; needle nozzle lor, 34 

P^iioles, Compaiiia de, compressor at, 38 

Performance of air compressors, 137-168 

Perry "The Steam Engine," 54 

Peterborough, Ont., hydraulic compressor at, 218 

Phila. and Reading C. and I. Co., compressed-air locomotives, 459 

Physical condition of machine drills, effects of, 299 

Pick machines for coal mining: Hardy, 371; Harrison, 366, 367; Ingersoll-Rand, 
369; Sullivan, 366, 368 

Pipe-line, for compressed-air haulage, 447-449; calculations for, 448; for com- 
pressed air transmission, 221-223, 231-234 

Pipe, nominal and actual diameters, 223, 230; bends in, 233, 234; joints in, 233; 
leakage, 231, 233, 449, 463; precautions in laying, 233, 449 

Piston air valve, 136 

Piston clearance in air-compressor cylinders, 54-71, 79-82 

Piston clearance in air engines, 243 

Piston inlet valves, Ingersoll-Rand, 107-109, 118 

Piston speed of compressors, 32, 73, 77, 86, 95, loi, 123 

Pitot tube for measuring compressed air, 475 

Plate valves, 109-118 



INDEX 501 

Plymouth Cordage Co., compressed-air locomotive at, 433 

Pneumatic displacement pumps, 409-415; consumption of air by, 411; Halsey, 

412; Latta-Martin, 411; Merrill, 410 
Pneumatic Engineering Co., 412 
Pneumelectric coal puncher, 290, 350, 372-374 
Pohle air-lift pump, 415 et seq. 
Poppet discharge valves, 119-1 23 
Poppet inlet valves, 99, loi; inertia of, loi; lift, loi; spring resistance, loi, 103, 

104-105; sticking of , 106 
Port area for air cylinders, 100, loi, 103, 107, 113, 123 
Porter, H. K., Co., compressed-air locomotives, 434, 435, 437-442, 443-445, 446, 

448, 451, 458, 459-461, 462,-466 
Power cost of compressed air, 150 
Preheating for compressed-air pumps, 405, 406, 407 
Prescott steam pump, 396 
Pressure of air as influencing freezing, 255-257 
Pressure regulators, 175 et seq. 
Proportions of compressor cylinders, 25, 90-92, 95 
Protection from freezing of surface air piping, 257 
Pulsator for "Electric-air" drill, 290, 291 

Pumping by compressed air, 256, 394-408; classification of methods, 394 
Pumps, direct-acting, 394-398; compound pumps operated by air, 405-408; 

cylinder dimensions of simple pumps, 398; duty of, 402; efficiencies, 402, 403, 

404, 408; freezing of moisture in, 403, 404, 405, 407; horse-power of, 400-402; 

terminal and mean air pressure, 401 ; volume of air consumed, 399 
Pumps operated by direct action of compressed air, 409-431; air-lift pumps, 

415-431; classification, 409; efficiencies, 419, 420, 421, 426, 431; pneumatic 

displacement pumps, 409-415 
"Punchers " (see Coal punchers) 

R 

"R" value of, 50 
Radialaxe coal cutter, 371 
Railway and Engineering Review, 210 
Ram-track channeler, Ingersoll-Rand Co., 382, 384 

Rand compressor, 2, 97; drill, 2; mechanical valve gear, 99; reheater, 264, 265 
Rand Mines Power Supply Co., air supplied by meter, 300, 301 
Randall, B. M., experiments by, on air-lift pump, 418 
Randolph, B. S., on compressed-air haulage plant, 465 
Rateau turbo-compressor, 42 
Rating of compressors, 137, 138 
Rawhide pinions for geared compressors, 38 
Read, T. T., on Anaconda compressed-air hoists, 249 
Receiver aftercoolers, 95, 170, 173; temperature of cooling water, 97, 174 
Receivers, air, 96, 169-174; baffle-plates, for 173; explosions in, 198-208; func- 
tions of, 169, 170; sizes of, 169; underground receivers, 171, 172, 256, 403 
Reciprocating coal cutters, 365 
Reciprocating drills, 273 et seq. 



602 INDEX 

Recording thermometer for compressors, 202, 207 

Records of work: hammer drills, 348; reciprocating drills, 309 

Reducing valves for compressed-air pumps, 256, 403 

Reduction of pressure as influencing deposition of moisture, 256-259 

Regulators, air-pressure, 175-190 

Reheaters for compressed-air engines, 245, 248, 251, 26:2-265, 4o7> 4o8; for 
channelers, 388, 390; IngersoU-Rand reheater, 263; internally-fired, 264; 
Sullivan reheater, 265; for underground work, 266; wet versus dry reheating, 
268 

Reheating compressed air, 245, 258-268; results of, 259; temperatures, 259-262; 
thermal cost of, 258, 259, 261 

Reheating temperatures, 259-262; reheating with steam, 267, 268 

Repairs of machine drills, 304; costs of, 305-307 

Resistance due to bends in air piping, 234 

"Return air" system of transmission, 246, 247; applied to pumping, 247, 413-415 

Richards, Frank, 90, 116, 145, 174, 246, 258, 267, 398 

Richard's formula for compressed-air pipe lines, 230 

Riedler compressors, 8, 20, 132; "express" discharge valve, 122; mechanically- 
controlled valves, 132-136 

Riedler, reheating experiments by, 262 

Risdon water-driven compressor, 29 

Risdon water wheel, 26 

Rix Compressor and Drill Co., 46 

Rix compressed-air locomotive, 458, 459 

Rix, E. A., on compressed-air pumps, 402, 404, 406, 411, 419 

Robey & Co., compressors, 8, 20, 41; plate valve, no 

Rock-drill testing, measuring tanks for, 470, 471, 478-485 

Rock drills (see Machine drills) 

Rock dust, cause of Silicosis, 316 

Rockford, III., tests on air-lift pump, 418 

Rose, A. H., on Victoria mine hydraulic compressor, 216 

Rose Deep mine, South Africa, rock drill tests, 300 

Rotary-bar coal-cutter, 359; Mavor & Coulson, 361, 362, 363 

Rotation devices for machine drills, 276-277, 278, 280, 318, 319, 321, 326 

"Rotator" hammer drills, Sullivan Machinery Co.: "air-tube," 329; "auger," 
329, 331; mountings for, 330; "water-tube," 230 

Rutland marble quarries, channeling at, 389 



Saunders, W. L., compressed-air data, 52 

Schneider & Co., Creusot, France, air-inlet valves, 99 

School of Mines Quarterly, 305 

Seeber, R. R., on Anaconda hoists, 249 

Sergeant drill, 274-277, 292 

Seymour, L. I., experiments on rock drills, 300 

Shaw Pneumatic Tool Co., hammer drill, 349 

Shaw, S. F., 431 



INDEX 503 

"Silent-chain" drive for compressors, 37, 40 

Silicosis ("miner's consumption"), 316 

Single-stage compression, work of, 54-60, 63-66 

Simple air-driven pumps, 396-398 

Siskol drill, 287 

"Skip- valve" for Nor walk high-pressure air cylinder, ic6 

Slimes and sands pumped by air-lift, 424-426 

"Slugger" drill. Chic. Pneum. Tool Co., 287 

Soap and water for cleaning air cylinders, 206 

Sommeiller, i; compressor, 2 

South African Association of Engineers, Trans, of, 300 

Spalding, C. M., on compressor explosions, 198 

Specific heat of air: at constant pressure and volume, 53 

Speed governorSjT 75-190 

Speer, F. W., tests on Victoria mine hydraulic compressor, 216 

Spiral- weld steel air pipe, 233 

Spool-valve drills, 274-282, 286-288 

Spool versus tappet valves for machine drills, 303 

Spray injection for compressors, 74, 75, 83, 85 

Spring-controlled air-valves, 101-107, 1 19-122 

Spring for inlet valves, loi; resistance of, loi, 103-105 

Stage compression, theory, 60 et seq., 86-98; at high altitudes, 196 

Stage compressors, 8, 13-19, 22, 35, 36-39, 41, 86-98; air card of, deductions 

from, 97; for high altitudes, 196; ratio of cylinder volumes, 90-92; work of, 

141 
Stanley heading machine, 375, 376 
Steam valves for compressors, 21, 23-25 
Steam versus compressed-air transmission of power, 3 
Steam, use underground, disadvantages of, 3 
Steam-driven, direct-acting pumps, 394-396 
Stephens & Son, Carn Brea, 281; "Climax Imperial" hammer drill, 318, 319; 

reciprocating drill, 280, 281 
St. Gothard tunnel, compressors at, 74 
"Stopehamer" drills, Ingersoll-Rand Co., 343-346 
"Stoper" drills (with air-feed standard), 339-347 
Storage-battery electric locomotives, 432 
Storage capacity of air locomotives, 434, 435, 436, 440, 441, 453, 458 et seq.; of 

pipe-lines, 448, 449 
Straight-line compressors, 8-14, 20, 28, 35, 36 
Stroke, length of, for compressors, 25 

Submergence of air-lift pumps, 417, 418, 419, 421, 426, 431 
Suction valves (see Inlet valves) 

Sullivan channelers, 381, 383, 388, 391; foot-pieces for air-lifts, 423 
Sullivan coal-cutters: "Iron-clad" chain machines, 351, 357, 359-361; coal-pick, 

366, 368, 372 
Sullivan compressors, 12, 14, 34, 35, 37, 99, 179; governors and unloaders, 178, 179, 

180; intercooler, 96; plate valve, no, 115; reheater, 264; air valve motions, 

99, 129, 131 



504 INDEX 

Sullivan drills: Air-feed hammer drills, 330, 332, 333; "Air-tube Rotator," 328; 
"Auger Rotator," 329, 331; "DR-6" hammer drill, 316, 317; "Hy-speed," 
288, 289; "Liteweight," 277-279, 312; mountings, 271, 332, 333; "Water- 
tube Rotator," 329 

Summers, L. L., experiments on air consumption of rock drills, 299 

Surface air piping, protection of, 233 

Susquehanna Coal Co., No. 6 Colliery, compressed-air haulage at, 462, 463 



Tailings pumps and wheels replaced by air-lift pumps, 424-426 

Tanks, capacity of, for compressed-air locomotives, 434, 435, 436, 440, 441, 453, 

458 et seq.; for measuring compressed air, 470, 471, 478-485 
Tappet drills, 288-290 

Tappet verstis spool valves for machine drills, 303 
Taylor, Chas. H., 210; hydraulic air compressor, 209 et seq., 267 
Technical Society of the Pacific Coast, Proceedings of, 402, 411, 418 
Temperatures employed in reheating, 245, 248, 251, 252, 257, 259-262, 266, 268 
Temperatures of compression, 5^, 44, 49, 51, 52, 73, 74, 86, 144, 145, 199, 200, 

202, 203, 204; of inlet air, ;^;^, 117, 144 
Temperatures of expansion, 238 

Temperatures, rate of increase of, in compression, 50, 51 
Temple-Ingersoll "electric-air" drill, 5, 290-292 
Tennessee Copper Co., compressor of, 21, 149 
Tests on compressors, 21, 45, 149-168 

Tests on machine drills, 299-301; for air consumption, 470, 4/1, 478-485 
Theoretical horse-power required to compress air, 54 et seq.; 138 et seq. 
Thermal cost of reheating, 258-259 
Thermal units, English and French, 51 
Thermodynamic laws, 47 et seq. 

"Thin-plate" valves: for discharge, 122; for inlet, 99 
Thomson-Houston turbo-compressor, 44 
Three-stage compression, work of, 141 
Three-stage compressors^ 86, 453-455 
Time-pressure curve for recording air measurements, 482 
Tiro General mine, Mex., air-lift pumping at, 431 
"Tool-om-eter," for measuring com.pressed air, 476-477 
Track channeler, Sullivan Machinery Co., 381 
Track resistance of mine cars, 452 
Transmission losses, comparison of air and steam, 3; in pipes, 221-234; loss of 

power, 221; loss of pressure, 222 cl seq. 
Transmission of power by compressed air, 221-234; at high pressures, 255 
Tripod mounting for drills, 270, 271 
"Triumph" drill, 293-294 

"Two-pipe" system of compressed-air transmission, 245-247 
Two-stage compression, work of, 60-64, 66-71, 141 
Turbo-compressors, 42-45; field of use, 44 



INDEX 505 



U 



Undercutting machines for coal, 350 et seq. 

Underground air receivers, 17!) 172, 256 

Underground compressed-air hoists, 248 

Underground reheaters, 266 

Unloaders for air cylinders, 182-^190; Allis-Chalmers, 185; Ingersoll-Rand, 182-185, 

190; Nordberg, 185-189; Rand "Imperial," 184; Sullivan, 179 
Unwin, W. C, on loss of pressure in pipe lines, 231; on reheating, 262 



Valve, adjustable cutoflf, for steam, 24 

Valveless drills, 293, 294, 320, 321, 327, 328, 336, 337 

Valves, air-inlet, 99-118, 124-136; Allis-Chalmers, 99, 129; area of, 100; Bailey 
& Co., 136; balanced piston valve, 24; chattering of, loi, 103; clack, 99; 
Corliss, 99, 107, 125, 127, 130, 131; Guttermuth, no; Humboldt, 72; inertia 
of, loi; Ingersoll-Rand "Hurricane inlet," loi, 107; Koster piston valve, 
136; Laidlaw-Dunn-Gordon, 99, loi, 102, 113, 127; Leyner flat annular 
valve, no; mechanically-controlled, 99; mushroom, loi, 107; Nordberg, 
99, 127, 185-187; Norwalk, 99, 102, 106, 125; requisites of, 99; resistance of, 
loi, 103-105; Riedler double-seat poppet, 132-136; "skip" valve, 106; 
springs for, 101-106; sticking of, 106; Sullivan, 99, 131 

Valves, air-delivery, 119-123, 124-136; air-cataract, 121; Allis-Chalmers, 129; 
area of, 122, 123; Corliss, 121, 124, 125; Ingersoll-Rand, 120; Laidlaw-Dunn- 
Gordon "Cincinnati" valve gear, 129; poppet, 119, 127, 129; Nordberg, 127; 
Norwalk, 121, 125; oil cataract, 121; plate valves, 122; Riedler double-seat 
poppet, 132-136; Riedler "Express," 121; Sullivan, 129, 131 

Valves for steam and of compressors, 21, 23-25 

Valves of machine drills, 274 et seq.; 313 et seq. (see under individual drills) 

Velocity of flow of air in pipes, 231, 232 

Vertical compressors, 8, 38,41 

Victoria copper mine, Mich., hydraulic air compressor, at 215, 220 

Victoria Falls and Transvaal Power Co., 42; turbo-compressors of, 42-43 

Village Main Reef mine. So. Africa, drill tests, 301 

Volumes and pressures of air, at altitudes above sea-level, 194; influence of 
reheating on, 258-262 

Volumetric output of compressors, 63, 79, 80, 100, 104-105 

Vulcan Iron Works, compressor, 46 

W 

"Wabbling-disk" meter, for compressed air, 477 
Wainwright water heater employed as reheater, 407 
Walker, G., Blake, 43, 44 
Walker Bros., compressors, 20; air valve, no 
Wandsworth, England, test on air-lift pump, 421 



506 INDEX 

Water attachments and tanks for rock drills, 278, 279, 283, 312, 315, 318, 319, 

322, 325, 328, 329, 346 
"Water" drills, Leyner-Ingersoll, 286, 313-316 
Water-driven compressors: at Alaska-Treadwell mine, 26; at Goleta mine, 29; 

at Morning mine, 32-33 
Water-jackets for compressor cylinders, 42, 59, 76-79, 97, 200, 207 
Water meter, modified for measuring flow of air, 477 
"Water-tube Rotator" drill, Sullivan Machinery Co., 329, 330 
Water-wheels for driving compressors: Knight, 26; Pelton, 26; peripheral velocity 

of, 31; Ridson, 26 
Waugh hammer drills, Denver Rock Drill Machinery Co.: "Clipper," 335, 336; 

"Dreadnaught," 337; stopers, 340-342 
Webb, R. L., compressor tests, 1526/ seq. 
Weber, F. C, on work of air-engines, 244 
Weeks, W. S., 478 

Weight and volume of dry air, 48, 469 
Weisbach, Mechanics of Engineering, 467 
Westinghouse Electric Co., turbo-compressor, 42, 43, 45 
Weston, E. M., book on "Rock Drills," 305 
Wet compressors, 72 e/ seq.; objections to, 84 
"Wet" reheating, 267, 268; versus dry reheating, 268 
Wet versus dry compressors, 82 
Weymouth, C. R., on reheating, 268 

Weymouth and Freeman, apparatus for measuring low-pressure air, 469-471 
Whitcomb, G. D. Co., Harrison coal-pick, 366, 367; hammer drill, 349 
Wickes Machinery Co., Murphy drill, 289 
Wilson colliery. Pa., compressed-air haulage at, 461 
Wilson, tests on air-lift pumps, 424 
"Wonder" hammer drill, 320, 321 

Wood drills: hammer type, 339; reciprocating, 279, 280 
Woodbridge, D. E., on Victoria mine hydraulic compressor, 216 
Work of compressors, with clearance, 63-71; without clearance, 54-63; of stage 

compressors, 60-63, 66-71, 97, 141 
Work done by air engines, 239-242 
Work done by air-lift pumps, 420, 421, 422, 426 
Work gained by reheating, 258, 259, 261, 262, 268 

Work required to compress air, 54-71, 83, 138-143; in stage compression, 97-141 
Worthington compound pump at Gwin mine, 407 
Worthington, Henry R., direct-acting pump, 394 
Worthington Pump and Machine Co., compressors, 46; water meter, 477 

Z 

Zahner, "Transmission of Power by Compressed Air," 74, 75 



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